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Abstract:

A railroad car truck for a railroad freight car, such as an autorack car,
has a bolster mounted cross-wise between two sideframes. The bolster ends
are mounted on respective spring groups carried by the sideframes. The
bolster can translate laterally relative to the sideframes. The side
frames are mounted to swing laterally relative to the wheel sets, and
hence relative to the rails. Resistance to lateral deflection is provided
by the resistance of the sideframes to the pendulum swinging motion, and
by shear in the spring groups. The truck has a doubled damper arrangement
of dampers in a four-cornered layout at each end of the bolster, giving a
flexing resistance to yaw between the sideframes and the bolster ends.
The doubled damper arrangement works against large wear plates mounted on
the sideframe columns. The large wear plates are mounted normal to the
dampers and square to the sideframes.

Claims:

1. A railroad freight car truck having a load rating, said truck
comprising: a bolster, sideframes, spring groups and wheelsets; said
bolster being mounted cross-wise to said sideframes; said bolster having
respective ends supported on respective ones of said spring groups
carried by said sideframes, said spring groups having vertical spring
rates; said sideframes being swingingly mounted on said wheelsets; said
bolster being moveable through a lateral displacement relative to said
sideframes, said lateral displacement having an overall magnitude and
including a first component associated with a first lateral stiffness,
kpendulum, opposing cross-wise swinging deflection of said
sideframes and a second component associated with a second lateral
stiffness, kspring shear, opposing sideways shear of said spring
groups; said first lateral stiffness being softer than said second
lateral stiffness; said bolster being movable in non-trivial yaw relative
to said sideframes; said truck having yaw resisting members mounted
yieldingly to oppose yawing of said bolster relative to said sideframes;
dampers mounted to work between said respective ends of said bolster and
the sideframes, said dampers having damper wedges; said damper wedges
each having a first face for working against an associated wear surface
in a friction relationship, and a second face for seating in a damper
wedge pocket, said first and second faces being angled with respect to
one another at a primary damper wedge angle, said primary damper wedge
angle being at least 35 degrees.

2. The railroad freight car truck of claim 1 wherein: said bolster has a
first end and a second end; four of said damper wedges are mounted at
said first end of said bolster, and four of said damper wedges are
mounted at said second end of said bolster; and said damper wedges also
have secondary damper wedge angles oriented cross-wise to said primary
damper wedge angles.

6. The railroad freight car truck of claim 1 wherein: said bolster has a
first end and a second end; four of said damper wedges are mounted at
said first end of said bolster, and four of said damper wedges are
mounted at said second end of said bolster; and said truck is free of (a)
a transom; (b) a frame brace; and (c) unsprung lateral bracing rods,
mounted between sideframes.

Description:

[0001] This application is a continuation of U.S. patent application Ser.
No. 12/582,368, filed Oct. 20, 2009, issued Sep. 6, 2011 as U.S. Pat. No.
8,011,306, which is a continuation of U.S. patent application Ser. No.
11/747,950, filed May 14, 2007, and issued Oct. 20, 2009 as U.S. Pat. No.
7,603,954, which is a continuation of U.S. patent application Ser. No.
11/363,520, filed Feb. 28, 2006, and issued Sep. 4, 2007 as U.S. Pat. No.
7,263,931, which is a divisional of U.S. patent application Ser. No.
10/355,374, filed Jan. 31, 2003, and issued on Feb. 28, 2006 as U.S. Pat.
No. 7,004,079, which is a continuation-in-part of U.S. patent application
Ser. No. 09/920,437, filed on Aug. 1, 2001, now U.S. Pat. No. 6,659,016;
and a continuation-in-part of U.S. patent application Ser. No.
10/210,797, filed Aug. 1, 2002, now U.S. Pat. No. 6,895,866; and a
continuation-in-part of U.S. patent application Ser. No. 10/210,853 also
filed Aug. 1, 2002, now U.S. Pat. No. 7,255,048. The specifications of
U.S. patent application Ser. Nos. 12/582,368, 11/747,950, 11/363,520 and
10/355,374 are being incorporated herein by reference.

FIELD OF THE INVENTION

[0002] This invention relates generally to rail road freight cars and to
trucks for use with rail road freight cars.

BACKGROUND OF THE INVENTION

[0003] Auto rack rail road cars are used to transport automobiles.
Typically, auto-rack rail road cars are loaded in the "circus loading"
manner, by driving vehicles into the cars from one end, and securing them
in place with chocks, chains or straps. When the trip is completed, the
chocks are removed, and the cars are driven out. The development of
autorack rail road cars can be traced back 80 or 90 years, when mass
production led to a need to transport large numbers of automobiles from
the factory to market.

[0004] Automobiles are a high value, relatively low density, relatively
fragile type of lading. Damage to lading due to dynamic loading in the
railcar may tend to arise principally in two ways. First, there are
longitudinal input loads transmitted through the draft gear due to train
line action or shunting. Second, there are vertical, rocking and
transverse dynamic responses of the rail road car to track perturbations
as transmitted through the rail car suspension. It would be desirable to
improve ride quality to lessen the chance of damage occurring.

[0005] In the context of longitudinal train line action, damage most often
occurs from two sources (a) slack run-in and run out; (b) humping or flat
switching. Rail road car draft gear have been designed against slack
run-out and slack run-in during train operation, and also against the
impact as cars are coupled together. Historically, common types of draft
gear, such as that complying with, for example, AAR specification
M-901-G, have been rated to withstand an impact at 5 m.p.h. (8 km/h) at a
coupler force of 500,000 Lbs. (roughly 2.2×106 N). Typically,
these draft gear have a travel of 23/4 to 31/4 inches in buff before
reaching the 500,000 Lbs. load, and before "going solid". The term "going
solid" refers to the point at which the draft gear exhibits a steep
increase in resistance to further displacement. If the impact is large
enough to make the draft gear "go solid" then the force transmitted, and
the corresponding acceleration imposed on the lading, increases sharply.
While this may be acceptable for ores, coal or grain, it is undesirably
severe for more sensitive lading, such as automobiles or auto parts,
rolls of paper, fresh fruit and vegetables and other high value consumer
goods such as household appliances or electronic equipment. Consequently,
from the relatively early days of the automobile industry there has been
a history of development of longer travel draft gear to provide lading
protection for relatively high value, low density lading, in particular
automobiles and auto parts, but also farm machinery, or tractors, or
highway trailers.

[0006] The subject of slack action is discussed at length in my co-pending
U.S. patent application Ser. No. 09/920,437 filed Aug. 1, 2001, now U.S.
Pat. No. 6,659,016, and incorporated herein by reference.

[0007] Since automobiles tend to be a relatively low density form of
lading as compared to grain, ores, or coal, the volumetric capacity of
the cars tends to be filled up before the weight of the reaches the
maximum allowable weight for the trucks. This has led to efforts to
increase the volumetric capacity of the cars. Over time, particularly in
the period of 1945-1970, autorack cars grew longer and taller. At
present, an autorack car may be up to about 90 feet long and 20 ft-2
inches tall. Autorack cars may typically have a tall, somewhat barn-like
housing. The housing has end doors that are intended to keep out thieves
and vandals.

[0008] The desire to increase the internal volume of the autorack car, and
the relatively light weight of the lading, led to the development of a
special 70 Ton rail road car truck for use with autorack cars. A 70 Ton
"special" truck is shown in the 1997 Car and Locomotive Cyclopedia
(Simmons-Boardman, Omaha, 1997) at page 726. The illustration indicates
that the total loading of the spring groups at solid is indicated as
70,166 Lbs. per spring group, giving a total of 140,334 Lbs. per truck
and 280,668 Lbs. per single unit autorack car. The spring rate is
indicated as 18,447 Lbs./in., per spring group or 36,894 Lbs./in for the
truck overall (there being one spring group per side frame, and two
spring groups per truck). The truck shown in the 1997 Cyclopedia is a
swing motion truck manufactured by National Castings Inc. In contrast to
a regular 70 Ton truck that has, typically, 33 inch diameter wheels, the
70 Ton special autorack truck has wheels that have a diameter of only 28
inches. This tends to allow for lower main deck wheel trackways, and
hence greater inside clearance height. In part, the use of such a truck
in an autorack car may reflect the low density of the lading. That is, a
regular 70 Ton truck is designed to carry a gross weight on rail of
110,000 Lbs, for a total full car weight of 220,000 Lbs. If the dead
sprung weight of a conventional single unit autorack car is 75-85,000
Lbs., and the unsprung weight is about 15,000 Lbs, that would leave about
120,000 Lbs., for lading. Assuming that a typical passenger sedan weighs
about 2500 Lbs., that would allow for about 48 automobiles before the
gross weight on rail would be exceeded. Even for larger, heavier
vehicles, weighing perhaps as much as 5000 Lbs., this would still give
some 24 light trucks, vans, or "sport utility vehicles". But the
volumetric capacity of a single unit autorack rail road car may be about
12-15 family sedans and perhaps fewer light trucks, vans, or SUV's. Thus
the autorack rail road car truck loading may often tend to be
significantly less than 110,000 lbs.

[0009] In contrast to the philosophy underlying the design of the 70 Ton
special 28 inch truck, the present inventor believes that it is
advantageous to use a truck having wheels larger than 33 inches in
diameter for auto rack rail road cars. Wheel life and maintenance are
dependent on wheel loading, and, for the same loading history, inversely
dependent on wheel diameter. A larger wheel may tend to have lower
operating stresses for the same lading; may tend to have a greater wear
allowance for braking; may tend to undergo fewer rotations than a wheel
of smaller diameter for the same distance traveled, and therefore may
tend to accumulate fewer cycles in terms of fatigue life; and may tend
not to get as hot during braking. All of these factors may tend to
increase wheel life and reduce maintenance. Further, a larger wheel
diameter may be used in conjunction with the use of longer springs. The
use of longer springs may permit the employment of springs having a
softer spring rate, giving a gentler ride. In terms of fatigue life and
wear, this in turn may tend to give a load history with reduced peak
loads, and lower frequency of those peak loads. Attainment of any one of
these advantages would be desirable.

[0010] In terms of dynamic response through the trucks, there are a number
of loading conditions to consider. First, there is a direct vertical
response in the "vertical bounce" condition. This may typically arise
when there is a track perturbation in both rails at the same point, such
as at a level crossing or at a bridge or tunnel entrance where there may
be a relatively sharp discontinuity in track stiffness. A second
"rocking" loading condition occurs when there are alternating track
perturbations, typically such as used formerly to occur with staggered
spacing of 39 ft rails. This phenomenon is less frequent given the
widespread use of continuously welded rails, and the generally lower
speeds, and hence lower dynamic forces, used for the remaining non-welded
track. A third loading condition arises from elevational changes between
the tracks, such as when entering curves in which case a truck may have a
tendency to warp. A fourth loading condition arises from truck "hunting",
typically at higher speeds, where the truck oscillates transversely
between the rails. During hunting, the trucks tend most often to deform
in a parallelogram manner. Fifth, lateral perturbations in the rails
sometimes arise where the rails widen or narrow slightly, or one rail is
more worn than another, and so on.

[0011] There are both geometric and historic factors to consider related
to these loading conditions and the dynamic response of the truck. One
historic factor is the near universal usage of the three-piece style of
freight car truck in North America. While other types of truck are known,
the three piece truck is overwhelmingly dominant in freight service in
North America. The three piece truck relies on a primary suspension in
the form of a set of springs trapped in a "basket" between the truck
bolster and the side frames. Rather than requiring independent suspension
of each wheel, for wheel load equalization a three piece truck uses one
set of springs, and the side frames pivot about the truck bolster ends in
a manner like a walking beam. It is a remarkably simple and durable
layout. However, the dynamic performance of the truck flows from that
layout. The 1980 Car & Locomotive Cyclopedia, states at page 669 that the
three piece truck offers "interchangeability, structural reliability and
low first cost but does so at the price of mediocre ride quality and high
cost in terms of car and track maintenance". It would be desirable to
retain many or all of these advantages while providing improved ride
quality.

[0012] In terms of rail road car truck suspension loading regimes, the
first consideration is the natural frequency of the vertical bounce
response. The static deflection from light car (empty) to maximum laded
gross weight (full) of a rail car at the coupler tends to be typically
about 2 inches. In addition, rail road car suspensions have a dynamic
range in operation, including a reserve travel allowance.

[0013] In typical historical use, springs were chosen to suit the
deflection under load of a full coal car, or a full grain car, or fully
loaded general purpose flat car. In each case, the design lading tended
to be very heavy relative to the rail car weight. For example, the live
load for a 286,000 lbs. car may be of the order of five times the weight
of the dead sprung load (i.e., the weight of the car, including truck
bolsters but less side frames, axles and wheels). Further, in these
instances, the lading may not be particularly sensitive to abusive
handling. That is, neither coal nor grain tends to be badly damaged by
poor ride quality. As a result, these cars tend to have very stiff
suspensions, with a dominant natural frequency in vertical bounce mode of
about 2 Hz. when loaded, and about 4 to 6 Hz. when empty. Historically,
much effort has been devoted to making freight cars light for at least
two reasons. First, the weight to be back hauled empty is kept low,
reducing the fuel cost of the backhaul. Second, as the ratio of lading to
car weight increases, a higher proportion of hauling effort goes into
hauling lading, rather than hauling the railcar.

[0014] By contrast, an autorack car, or other type of car for carrying
relatively high value, low density lading such as auto parts, electronic
consumer goods, or white goods more generally, has the opposite loading
profile. A two unit articulated autorack car may have a light car (i.e.,
empty) weight of 165,000 lbs., and a lading weight when fully loaded of
only 35-40,000 lbs., per car body unit. That is, not only may the weight
of the lading be less than the sprung weight of the rail road car unit,
it may be less than 40% of the car weight. The lading typically has a
high, or very high, ratio of value to weight. Unlike coal or grain,
automobiles are relatively fragile, and hence more sensitive to a gentle
(or a not so gentle) ride. As a relatively fragile, high value, high
revenue form of lading, it may be desirable to obtain superior ride
quality to that suitable for coal or grain.

[0015] Historically, auto rack cars were made by building a rack structure
on top of a general purpose flat car. As such, the resultant car was
sprung for the flat car design loads. When loaded with automobiles, this
might yield a vertical bounce natural frequency in the range of 3 Hz. It
would be preferable for the railcar vertical bounce natural frequency to
be on the order of 1.4 Hz or less when loaded. Since this natural
frequency varies as the square root of the quotient obtained by dividing
the spring rate of the suspension by the overall sprung mass, it is
desirable to reduce the spring constant, to increase the mass, or both.

[0016] One way to improve ride quality is to increase the dead sprung
weight of the rail road car body. Deliberately increasing the mass of a
freight car is counter intuitive, since many years of effort has gone
into reducing the weight of rail cars relative to the weight of the
lading for the reasons noted above. One manufacturer, for example,
advertises a light weight aluminum auto-rack car. However, given the high
value and low density of the lading, adding weight may be reasonable to
obtain a desired level of ride quality. Further, auto rack rail cars tend
to be tall, long, and thin, with the upper deck loads carried at a
relatively high location as measured from top of rail. A significant
addition of weight at a low height relative to top of rail may also be
beneficial in reducing the height of the center of gravity of the loaded
car.

[0017] Another way to improve ride quality is to decrease the spring rate.
Decreasing the spring rate involves further considerations. Historically
the deck height of a flat car tended to be very closely related to the
height of the upper flange of the center sill. This height was itself
established by the height of the cap of the draft pocket. The size of the
draft pocket was standardized on the basis of the coupler chosen, and the
allowable heights for the coupler knuckle. The deck height usually worked
out to about 41 inches above top of rail. For some time auto rack cars
were designed to a 19 ft height limit. To maximize the internal loading
space, it has been considered desirable to lower the main deck as far as
possible, particularly in tri-level cars. Since the lading is relatively
light, the rail car trucks have tended to be light as well, such as 70
Ton trucks, as opposed to 100, 110 or 125 Ton trucks for coal, ore, or
grain cars at 263,000, 286,000 or 315,000 lbs. gross weight on rail.
Since the American Association of Railroads (AAR) specifies a minimum
clearance of 5'' above the wheels, the combination of low deck height,
deck clearance, and minimum wheel height set an effective upper limit on
the spring travel, and reserve spring travel range available. If softer
springs are used, the remaining room for spring travel below the decks
may well not be sufficient to provide the desired reserve height. In
consequence, the present inventor proposes, contrary to lowering the main
deck, that the main deck be higher than 42 inches to allow for more
spring travel.

[0018] As noted above, many previous auto rack cars have been built to a
19 ft height. Another major trend in recent years has been the advent of
"double stack" intermodal container cars capable of carrying two shipping
containers stacked one above the other in a well or to other freight cars
falling within the 20 ft 2 in. height limit of AAR plate H. Many main
lines have track clearance profiles that can accommodate double stack
cars. Consequently, it is now possible to use auto rack cars built to the
higher profile of the double stack intermodal container cars.

[0019] While decreasing the primary vertical bounce natural frequency
appears to be advantageous for auto rack rail road cars generally,
including single car unit auto rack rail road cars, articulated auto rack
cars may also benefit not only from adding ballast, but from adding
ballast preferentially to the end units near the coupler end trucks. As
explained more fully in the description below, the interior trucks of
articulated cars tend to be more heavily burdened than the end trucks,
primarily because the interior trucks share loads from two adjacent car
units, while the coupler end trucks only carry loads from one end of one
car unit. It would be advantageous to even out this loading so that the
trucks have roughly similar vertical bounce frequencies.

[0020] Three piece trucks currently in use tend to use friction dampers,
sometimes assisted by hydraulic dampers such as can be mounted, for
example, in the spring set. Friction damping has most typically been
provided by using spring loaded blocks, or snubbers, mounted with the
spring set, with the friction surface bearing against a mating friction
surface of the columns of the side frames, or, if the snubber is mounted
to the side frame, then the friction surface is mounted on the face of
the truck bolster. There are a number of ways to do this. In some
instances, as shown at p. 847 of the 1961 Car Builders Cyclopedia lateral
springs are housed in the end of the truck bolster, the lateral springs
pushing horizontally outward on steel shoes that bear on the vertical
faces of the side columns of the side frames. This provides roughly
constant friction (subject to the wear of the friction faces), without
regard to the degree of compression of the main springs of the
suspension.

[0021] In another approach, as shown at p. 715 of the 1997 Car &
Locomotive Cyclopedia, one of the forward springs in the main spring
group, and one of the rearward springs in the main spring group bear upon
the underside, or short side, of a wedge. One of the long sides,
typically an hypotenuse of a wedge, engages a notch, or seat, formed near
the outboard end of the truck bolster, and the third side has the
friction face that abuts, and bears against, the friction face of the
side column (either front or rear, as the case may be), of the side
frame. The action of this pair of wedges then provides damping of the
various truck motions. In this type of truck the friction force varies
directly with the compression of the springs, and increases and decreases
as the truck flexes. In the vertical bounce condition, both friction
surfaces work in the same direction. In the warping direction (when one
wheel rises or falls relative to the other wheel on the same side, thus
causing the side frame to pivot about the truck bolster) the friction
wedges work in opposite directions against the restoring force of the
springs.

[0022] The "hunting" phenomenon has been noted above. Hunting generally
occurs on tangent (i.e., straight) track as railcar speed increases. It
is desirable for the hunting threshold to occur at a speed that is above
the operating speed range of the rail car. During hunting the side frames
tend to want to rotate about a vertical axis, to a non-perpendicular
angular orientation relative to the truck bolster sometimes called
"parallelogramming" or lozenging. This will tend to cause angular
deflection of the spring group, and will tend to generate a squeezing
force on opposite diagonal sides of the wedges, causing them to tend to
bear against the side frame columns. This diagonal action will tend to
generate a restoring moment working against the angular deflection. The
moment arm of this restoring force is proportional to half the width of
the wedge, since half of the friction plate lies to either side of the
centerline of the side frame. This tends to be a relatively weak moment
connection, and the wedge, even if wider than normal, tends to be
positioned over a single spring in the spring group.

[0023] Typically, for a truck of fixed wheelbase length, there is a
trade-off between wheel load equalization and resistance to hunting.
Where a car is used for carrying high density commodities at low speeds,
there may tend to be a higher emphasis on maintaining wheel load
equalization. Where a car is light, and operates at high speed there will
be a greater emphasis on avoiding hunting. In general, the parallelogram
deformation of the truck in hunting may be deterred by making the truck
laterally more stiff. One approach to discouraging hunting is to use a
transom, typically in the form of a channel running from between the side
frames below the spring baskets. Another approach is to use a frame
brace.

[0024] One way to address the hunting issue is to employ a truck having a
longer wheelbase, or one whose length is proportionately great relative
to its width. For example, at present two axle truck wheelbases may range
from about 5'-3'' to 6'-0''. However, the standard North American track
gauge is 4'-81/2'', giving a wheelbase to track width ratio possibly as
small as 1.12. At 6'-0'' the ratio is roughly 1.27. It would be
preferable to employ a wheelbase having a longer aspect ratio relative to
the track gauge. As described herein, one aspect of the present invention
employs a truck with a longer wheelbase, which may be about 80 to 86
inches, giving a ratio of 1.42 or 1.52. This increase in wheelbase length
may tend also to be benign in terms of wheel loading equalization.

[0025] In a typical spring seat and spring group arrangement, the side
frame window may typically be of the order of 21 inches in height from
the spring seat base to the underside of the overarching compression
member, and the width of the side frame window between the wear plates on
the side frame columns is typically about 18'', giving a side frame
window that is taller than wide in the ratio of about 7:6. Similarly, the
bottom spring seat has a base that is typically about 18 inches long to
correspond to the width of the side frame window, and about 16 inches
wide in the transverse direction, that is being longer than wide. It may
be advantageous to make the side frame windows wider, and the spring seat
correspondingly longer to accommodate larger diameter long travel springs
with a softer spring rate or a larger number of softer coils of smaller
diameter. At the same time, lengthening the wheel base of the truck may
also be advantageous since it is thought that a longer wheelbase may
ameliorate truck hunting performance, as noted above. Such a design
change is counter-intuitive since it may generally be desired to keep
truck size small, and widening the unsupported window span may not have
been considered desirable heretofore.

[0026] Another way to raise the hunting threshold is to increase the
parallelogram stiffness between the bolster and the side frames. It is
possible, as described herein, to employ pairs of damper wedges, of
comparable size to those previously used, the two wedges being placed
side by side and each individually supported by a different spring, or
being the outer two wedges in a three deep spring group, to give a larger
moment arm to the restoring force and to the damping associated with that
force.

[0027] One determinant of overall ride quality is the dynamic response to
lateral perturbations. That is, when there is a lateral perturbation at
track level, the rigid steel wheelsets of the truck may be pushed
sideways relative to the car body. Lateral perturbations may arise for
example from uneven track, or from passing over switches or from turnouts
and other track geometry perturbations. When the train is moving at
speed, the time duration of the input pulse due to the perturbation may
be very short.

[0028] The suspension system of the truck reacts to the lateral
perturbation. It is generally desirable for the force transmission to be
relatively low. High force transmissibility, and corresponding high
lateral acceleration, may tend not to be advantageous for the lading.
This is particularly so if the lading includes relatively fragile goods,
such as automobiles, electronic equipment, white goods, and other
consumer products. In general, the lateral stiffness of the suspension
reflects the combined displacement of (a) the sideframe between (i) the
pedestal bearing adapter and (ii) the bottom spring seat (that is, the
sideframes swing laterally as a pendulum with the pedestal bearing
adapter being the top pivot point for the pendulum); and (b) the lateral
deflection of the springs between (i) the lower spring seat in the
sideframe and (ii) the upper spring mounting against the underside of the
truck bolster, and (c) the moment and the associated angular displacement
between the (i) spring seat in the sideframe and (ii) the upper spring
mounting against the underside of the truck bolster.

[0029] In a conventional rail road car truck, the lateral stiffness of the
spring groups is sometimes estimated as being approximately 1/2 of the
vertical spring stiffness. Thus the choice of vertical spring stiffness
may strongly affect the lateral stiffness of the suspension. The vertical
stiffness of the spring groups may tend to yield a vertical deflection at
the releasable coupler from the light car (i.e., empty) condition to the
fully laden condition of about 2 inches. For a conventional grain or coal
car subject to a 286,000 lbs., gross weight on rail limit, this may imply
a dead sprung load of some 50,000 lbs., and a live sprung load of some
220,000 lbs., yielding a spring stiffness of 25-30,000 lbs./in., per
spring group (there being, typically, two groups per truck, and two
trucks per car). This may yield a lateral spring stiffness of 13-16,000
lbs./in per spring group. It should be noted that the numerical values
given in this background discussion are approximations of ranges of
values, and are provided for the purposes of general order-of-magnitude
comparison, rather than as values of a specific truck.

[0030] The second component of stiffness relates to the lateral deflection
of the sideframe itself. In a conventional truck, the weight of the
sprung load can be idealized as a point load applied at the center of the
bottom spring seat. That load is carried by the sideframe to the pedestal
seat mounted on the bearing adapter. The vertical height difference
between these two points may be in the range of perhaps 12 to 18 inches,
depending on wheel size and sideframe geometry. For the general purposes
of this description, for a truck having 36 inch wheels, 15 inches (±)
might be taken as a roughly representative height.

[0031] The pedestal seat may typically have a flat surface that bears on
an upwardly crowned surface of the bearing adapter. The crown may
typically have a radius of curvature of about 60 inches, with the center
of curvature lying below the surface (i.e., the surface is concave
downward).

[0032] When a lateral shear force is imposed on the springs, there is a
reaction force in the bottom spring seat that will tend to deflect the
sideframe, somewhat like a pendulum. When the sideframe takes on an
angular deflection in one direction, the line of contact of the flat
surface of the pedestal seat with the crowned surface of the bearing
adapter will tend to move along the arc of the crown in the opposite
direction. That is, if the bottom spring seat moves outboard, the line of
contact will tend to move inboard. This motion is resisted by a moment
couple due to the sprung weight of the car on the bottom spring seat,
acting on a moment arm between (a) the line of action of gravity at the
spring seat and (b) the line of contact of the crown of the bearing
adapter. For a 286,000 lbs. car the apparent stiffness of the sideframe
may be of the order of 18,000-25,000 lbs./in, measured at the bottom
spring seat. That is, the lateral stiffness of the sideframe (i.e., the
pendulum action by itself) can be greater than the (already relatively
high) lateral stiffness of the spring group in shear, and this apparent
stiffness is proportional to the total sprung weight of the rail car
(including lading). When taken as being analogous to two springs in
series, the overall equivalent lateral spring stiffness may be of the
order of 8,000 lbs./in. to 10,000 lbs./in., per sideframe. A car designed
for lesser weights may have softer apparent stiffness. This level of
stiffness may not always yield as smooth a ride as may be desired.

[0033] There is another component of spring stiffness due to the unequal
compression of the inside and outside portions of the spring group as the
bottom spring seat rotates relative to the upper spring group mount under
the bolster. This stiffness, which is additive to (that is, in parallel
with) the stiffness of the sideframe, can be significant, and may be of
the order of 3000-3500 lbs./in per spring group, depending on the
stiffness of the springs and the layout of the group. Other second and
third order effects are neglected for the purpose of this description.
The total lateral stiffness for one sideframe, including the spring
stiffness, the pendulum stiffness and the spring moment stiffness, for a
S2HD 110 Ton truck may be about 9200 lbs/inch per side frame.

[0034] It has been observed that it may be preferable to have springs of a
given vertical stiffness to give certain vertical ride characteristics,
and a different characteristic for lateral perturbations. In particular,
a softer lateral response may be desired at high speed (greater than
about 50 m.p.h.) and relatively low amplitude to address a truck hunting
concern, while a different spring characteristic may be desirable to
address a low speed (roughly 10-25 m.p.h.) roll characteristic,
particularly since the overall suspension system may have a roll mode
resonance lying in the low speed regime.

[0035] An alternate type of three piece truck is the "swing motion" truck.
One example of a swing motion truck is shown at page 716 in the 1980 Car
and Locomotive Cyclopedia (1980, Simmons-Boardman, Omaha). This
illustration, with captions removed, is the basis of FIGS. 1a, 1b and 1c,
herein, labeled "Prior Art". Since the truck has both lateral and
longitudinal axes of symmetry, the artist has only shown half portions of
the major components of the truck. The particular example illustrated is
a swing motion truck produced by National Castings Inc., more commonly
referred to as "NACO". Another example of a NACO Swing Motion truck is
shown at page 726 of the 1997 Car and Locomotive Cyclopedia (1997,
Simmons-Boardroom, Omaha). An earlier swing motion three piece truck is
shown and described in U.S. Pat. No. 3,670,660 of Weber et al., issued
Jun. 20, 1972, the specification of which is incorporated herein by
reference.

[0036] In a swing motion truck, the sideframe is mounted as a "swing
hanger" and acts much like a pendulum. In contrast to the truck described
above, the bearing adapter has an upwardly concave rocker bearing
surface, having a radius of curvature of perhaps 10 inches and a center
of curvature lying above the bearing adapter. A pedestal rocker seat
nests in the upwardly concave surface, and has itself an upwardly concave
surface that engages the rocker bearing surface. The pedestal rocker seat
has a radius of curvature of perhaps 5 inches, again with the center of
curvature lying upwardly of the rocker.

[0037] In this instance, the rocker seat is in dynamic rolling contact
with the surface of the bearing adapter. The upper rocker assembly tends
to act more like a hinge than the shallow crown of the bearing adapter
described above. As such, the pendulum may tend to have a softer, perhaps
much softer, response than the analogous conventional sideframe.
Depending on the geometry of the rocker, this may yield a sideframe
resistance to lateral deflection in the order of 1/4 (or less) to about
1/2 of what might otherwise be typical. If combined in series with the
spring group stiffness, it can be seen that the relative softness of the
pendulum may tend to become the dominant factor. To some extent then, the
lateral stiffness of the truck becomes less strongly dependent on the
chosen vertical stiffness of the spring groups at least for small
displacements. Furthermore, by providing a rocking lower spring seat, the
swing motion truck may tend to reduce, or eliminate, the component of
lateral stiffness that may tend to arise because of unequal compression
of the inboard and outboard members of the spring groups when the
sideframe has an angular displacement, thus further softening the lateral
response.

[0038] In the truck of U.S. Pat. No. 3,670,660 the rocking of the lower
spring seat is limited to a range of about 3 degrees to either side of
center, and a transom extends between the sideframes, forming a rigid,
unsprung, lateral connecting member between the rocker plates of the two
sideframes. In this context, "unsprung" refers to the transom being
mounted to a portion of the truck that is not resiliently isolated from
the rails by the main spring groups.

[0039] When the three degree condition is reached, the rockers "lock-up"
against the side frames, and the dominant lateral displacement
characteristic is that of the main spring groups in shear, as illustrated
and described by Weber. The lateral, unsprung, sideframe connecting
member, namely the transom, has a stop that engages a downwardly
extending abutment on the bolster to limit lateral travel of the bolster
relative to the sideframes. This use of a lateral connecting member is
shown and described in U.S. Pat. No. 3,461,814 of Weber, issued Mar. 7,
1967, also incorporated herein by reference. As noted in U.S. Pat. No.
3,670,660 the use of a spring plank had been known, and the use of an
abutment at the level of the spring plank tended to permit the end of
travel reaction to the truck bolster to be transmitted from the
sideframes at a relatively low height, yielding a lower overturning
moment on the wheels than if the end-of-travel force were transmitted
through gibs on the truck bolster from the sideframe columns at a
relatively greater height. The use of a spring plank in this way was
considered advantageous.

[0040] In Canadian Patent 2,090,031, (issued Apr. 15, 1997 to Weber et
al.) noting the advent of lighter weight, low deck cars, Weber et al.,
replaced the transom with a lateral rod assembly to provide a rigid,
unsprung connection member between the platforms of the rockers of the
lower spring seats. As noted above, one type of car in which relative
lightness and a low main deck has tended to be found is an Autorack car.

[0041] For the purposes of rapid estimation of truck lateral stiffness,
the following formula can be used:

ktruck=2×[(ksideframe)-1+(kspring
shear)-1]-1

[0042] where

[0043] ksideframe=[kpendulum+kspring moment]

[0044] kspring shear=The lateral spring constant for the spring group
in shear.

[0045] kpendulum=The force required to deflect the pendulum per unit
of deflection, as measured at the center of the bottom spring seat.

[0046] kspring moment=The force required to deflect the bottom spring
seat per unit of sideways deflection against the twisting moment caused
by the unequal compression of the inboard and outboard springs.

[0047] For the range of motion that may typically be of interest, and for
small angles of deflection, kpendulum can be taken as being
approximately constant at, for example, the value obtained for deflection
of one degree. This may tend to be a sufficiently accurate approximation
for the purposes of general calculation.

[0048] In a pure pendulum, the lateral constant for small angles
approximates k=W/L, where k is the lateral constant, W is the weight, and
L is the pendulum length. Further, for the purpose of rapid comparison of
the lateral swinging of the sideframes, an equivalent pendulum length for
small angles of deflection can be defined as Leq=W/kpendulum.
In this equation W represents the sprung weight borne by that sideframe,
typically 1/4 of the total sprung weight for a symmetrical single unit
rail car. For a conventional truck Leq may be of the order of about
3 or 4 inches. For a swing motion truck, Leq may be of the order of
about 10 to 15 inches.

[0049] It is also possible to define the pendulum lateral stiffness (for
small angles) in terms of the length of the pendulum, the radius of
curvature of the rocker, and the design weight carried by the pendulum
according to the formula:

[0054] Lpendulum=the length of the pendulum, being the vertical
distance from the contact surface of the bearing adapter to the bottom
spring seat

[0055] Rcurvature=the radius of curvature of the rocker surface

[0056] Following from this, if the pendulum stiffness is taken in series
with the lateral spring stiffness, then the resultant overall lateral
stiffness can be obtained. Using this number in the denominator, and the
design weight in the numerator yields a length, effectively equivalent to
a pendulum length if the entire lateral stiffness came from an equivalent
pendulum according to

Lresultant=W/klateral total

[0057] For a conventional truck with a 60 inch radius of curvature rocker,
and stiff suspension, this length, Lresultant may be of the order of
6-8 inches, or thereabout.

[0058] So that the present invention may better be understood by
comparison, in the prior art illustration of FIGS. 1a, 1b and 1c, a NACO
swing motion truck is identified generally as A20. Inasmuch as the truck
is symmetrical about the truck center both from side-to-side and
lengthwise, the artist has shown only half of the bolster, identified as
A22, and half of one of the sideframes, identified as A24.

[0059] In the customary manner, sideframe A24 has defined in it a
generally rectangular window A26 that admits one of the ends of the
bolster A28. The top boundary of window A26 is defined by the sideframe
arch, or compression member identified as top chord member A30, and the
bottom of window A26 is defined by a tension member, identified as bottom
chord A32. The fore and aft vertical sides of window A26 are defined by
sideframe columns A34.

[0060] At the swept up ends of sideframe A24 there are sideframe pedestal
fittings A38 which each accommodate an upper rocker identified as a
pedestal rocker seat A40, that engages the upper surface of a bearing
adapter A42. Bearing adapter A42 itself engages a bearing mounted on one
of the axles of the truck adjacent one of the wheels. A rocker seat A40
is located in each of the fore and aft pedestals, the rocker seats being
longitudinally aligned such that the sideframe can swing transversely
relative to the rolling direction of the truck A20 generally in what is
referred to as a "swing hanger" arrangement.

[0061] The bottom chord of the sideframe includes pockets A44 in which a
pair of fore and aft lower rocker bearing seats A46 are mounted. The
lower rocker seat A48 has a pair of rounded, tapered ends or trunnions
A50 that sit in the lower rocker bearings A48, and a medial platform A52.
An array of four corner bosses A54 extend upwardly from platform A52.

[0062] An unsprung, lateral, rigid connecting member in the nature of a
spring plank, or transom A60 extends cross-wise between the sideframes in
a spaced apart, underslung, relationship below truck bolster A22. Transom
A60 has an end portion that has an array of four apertures A62 that pick
up on bosses A54. A grouping, or set of springs A64 seats on the end of
the transom, the corner springs of the set locating above bosses A54.

[0063] The spring group, or set A64, is captured between the distal end of
bolster A22 and the end portion of transom A60. Spring set A64 is placed
under compression by the weight of the rail car body and lading that
bears upon bolster A22 from above. In consequence of this loading, the
end portion of transom A60, and hence the spring set, are carried by
platform A54. The reaction force in the springs has a load path that is
carried through the bottom rocker A70 (made up of trunnions A50 and lower
rocker bearings A48) and into the sideframe A22 more generally.

[0064] Friction damping is provided by damping wedges A72 that seat in
mating bolster pockets A74. Bolster pockets A74 have inclined damper
seats A76. The vertical sliding faces of the friction damper wedges then
ride up an down on friction wear plates A80 mounted to the inwardly
facing surfaces of the sideframe columns.

[0065] The "swing motion" truck gets its name from the swinging motion of
the sideframe on the upper rockers when a lateral track perturbation is
imposed on the wheels. The reaction of the sideframes is to swing, rather
like pendula, on the upper rockers. When this occurs, the transom and the
truck bolster tend to shift sideways, with the bottom spring seat
platform rotating on the lower rocker.

[0066] The upper rockers are inserts, typically of a hardened material,
whose rocking, or engaging, surface A80 has a radius of curvature of
about 5 inches, with the center of curvature (when assembled) lying above
the upper rockers (i.e., the surface is upwardly concave).

[0067] As noted above, one of the features of a swing motion truck is that
while it may be quite stiff vertically, and while it may be resistant to
parallelogram deformation because of the unsprung lateral connection
member, it may at the same time tend to be laterally relatively soft.

[0068] The use of multiple variable friction force dampers in which the
wedges are mounted over members of the spring group, is shown in U.S.
Pat. No. 3,714,905 of Barber, issued Feb. 6, 1973. The damper arrangement
shown by Barber is not apparently presently available in the market, and
does not seem ever to have been made available commercially.

[0069] Notably, the damper wedges shown in Barber appear to have
relatively sharply angled wedges, with an included angle between the
friction face (i.e., the face bearing against the side frame column) and
the sliding face (i.e., the angled face seated in the damper pocket
formed in the bolster, typically the hypotenuse) of roughly 35 degrees.
The angle of the third, or opposite, horizontal side face, namely the
face that seats on top of the vertically oriented spring, is the
complementary angle, in this example, being about 55 degrees. It should
be noted that as the angle of the wedge becomes more acute, (i.e.,
decreasing from about 35 degrees) the wedge may have an undesirable
tendency to jam in the pocket, rather than slide.

[0070] Barber, above, shows a spring group of variously sized coils with
four relatively small corner coils loading the four relatively sharp
angled dampers. From the relative sizes of the springs illustrated, it
appears that Barber was contemplating a spring group of relatively
traditional capacity--a load of about 80,000 lbs., at a "solid" condition
of 3 1/16 inches of travel, for example, and an overall spring rate for
the group of about 25,000 lbs/inch, to give 2 inches of overall rail car
static deflection for about 200,000 lbs live load.

[0071] Apparently keeping roughly the same relative amount of damping
overall as for a single damper, Barber appears to employ individual B331
coils (k=538 lb/in, (±)) under each friction damper, rather than a
B432 coil (k=1030 lb/in, (±)) as might typically have been used under
a single damper for a spring group of the same capacity. As such, it
appears that Barber contemplated that springs accounting for somewhat
less than 15% of the overall spring group stiffness would underlie the
dampers.

[0072] These spring stiffnesses might typically be suitable for a rail
road car carrying iron ore, grain or coal, where the lading is not overly
fragile, and the design ratio of live load to dead sprung load is
typically greater than 3:1. It might not be advantageous for a rail road
car for transporting automobiles, auto parts, consumer electronics or
other white goods of relatively low density and high value where the
design ratio of live load to dead sprung load may be well less than 2:1,
and quite possibly lying in the range of 0.4:1 to 1:1.

[0073] In the past, spring groups have been arranged such that the spring
loading under the dampers has been proportionately small. That is, the
dampers have typically been seated on side spring coils, as shown in the
AAR standard spring groupings shown in the 1997 Car & Locomotive
Cyclopedia at pages 743-746, in which the side spring coils, inner and
outer as may be, are often B321, B331, B421, B422, B432, or B433 springs
as compared to the main spring coils, such that the springs under the
dampers have lower spring rates than the other coil combinations in the
other positions in the spring group. As such, the dampers may be driven
by less than 15% of the total spring stiffness of the group generally.

[0074] In U.S. Pat. No. 5,046,431 of Wagner, issued Sep. 10, 1991, the
standard inboard-and-outboard gib arrangement on the truck bolster was
replaced by a single central gib mounted on the side frame column for
engaging the shoulders of a vertical channel defined in the end of the
truck bolster. In doing this, the damper was split into inboard and
outboard portions, and, further, the inboard and outboard portions,
rather than lying in a common transverse vertical plane, were angled in
an outwardly splayed orientation.

[0075] Wagner's gib and damper arrangement may not necessarily be
desirable in obtaining a desired level of ride quality. In obtaining a
soft ride it may be desirable that the truck be relatively soft not only
in the vertical bounce direction, but also in the transverse direction,
such that lateral track perturbations can be taken up in the suspension,
rather than be transmitted to the car body, (and hence to the lading), as
may tend undesirably to happen when the gibs bottom out (i.e., come into
hard abutting contact with the side frame) at the limit of horizontal
travel.

[0076] The present inventor has found it desirable that there be an
allowance for lateral travel of the truck bolster relative to the wheels
of the order of 1 to 11/2 inches to either side of a neutral central
position. Wagner does not appear to have been concerned with this issue.
On the contrary, Wagner appears to show quite a tight gib clearance, with
relatively little travel before solid contact. Furthermore, transverse
displacement of the truck bolster relative to the side frame is typically
resiliently resisted by the horizontal shear in the spring groups, and by
the pendulum motion of the side frames rocking on the crowns of the
bearing adapters, these two components being combined like springs in
series. Wagner's canted dampers appear to make lateral translation of the
bolster stiffer, rather than softer. This may not be advantageous for
relatively fragile lading. In the view of the present inventor, while it
is advantageous to increase resistance to the hunting phenomenon, it may
not be advantageous to do so at the expense of increasing lateral
stiffness.

[0077] In the damper groups themselves, it is thought that parallelogram
deflection of the truck such that the truck bolster is not perpendicular
to the side frame, as during hunting, may tend to cause the dampers to
try to twist angularly in the damper seats. In that situation one corner
of the damper may tend to be squeezed more tightly than the other. As a
result, the tighter corner may try to retract relative to the less tight
corner, causing the damper wedge to squirm and rotate somewhat in the
pocket. This tendency to twist may also tend to reduce the squaring, or
restoring force that tends to move the truck back into a condition in
which the truck bolster is square relative to the side frames.

[0078] Consequently, it may be desirable to discourage this twisting
motion by limiting the freedom to twist, as, for example, by introducing
a groove or ridge, or keyway, or channel feature to govern the operation
of the spring in the damper pocket. It may also be advantageous to use a
split wedge to discourage twisting, such that one portion of the wedge
can move relative to the other, thus finding a different position in a
linear sense without necessarily forcing the other portion to twist.
Further still, it may be advantageous to employ a means for encouraging a
laterally inboard portion of the damper, or damper group, to be biased to
its most laterally inboard position, and a laterally outboard portion of
the damper, or the damper group, to be biased to its most laterally
outboard position. In that way, the moment arm of the restoring force may
tend to remain closer to its largest value. One way to do this, as
described in the description of the invention, below, is to add a
secondary angle to the wedge.

[0079] In the terminology herein, wedges have a primary angle ψ,
namely the included angle between (a) the sloped damper pocket face
mounted to the truck bolster, and (b) the side frame column face, as seen
looking from the end of the bolster toward the truck center. This is the
included angle described above. A secondary angle is defined in the plane
of angle ψ, namely a plane perpendicular to the vertical longitudinal
plane of the (undeflected) side frame, tilted from the vertical at the
primary angle. That is, this plane is parallel to the (undeflected) long
axis of the truck bolster, and taken as if sighting along the back side
(hypotenuse) of the damper.

[0080] The secondary angle β is defined as the lateral rake angle
seen when looking at the damper parallel to the plane of angle ψ. As
the suspension works in response to track perturbations, the wedge forces
acting on the secondary angle will tend to urge the damper either inboard
or outboard according to the angle chosen. Inasmuch as the tapered region
of the wedge may be quite thin in terms of vertical through-thickness, it
may be desirable to step the sliding face of the wedge (and the
co-operating face of the bolster seat) into two or more portions. This
may be particularly so if the angle of the wedge is large.

[0081] Split wedges and two part wedges having a chevron, or chevron like,
profile when seen in the view of the secondary angle can be used.
Historically, split wedges have been deployed as a pair over a single
spring, the split tending to permit the wedges to seat better, and to
remain better seated, under twisting condition than might otherwise be
the case. The chevron profile of a solid wedge may tend to have the same
intent of preventing rotation of the sliding face of the wedge relative
to the bolster in the plane of the primary angle of the wedge. Split
wedges and compound profile wedges can be employed in pairs as described
herein.

[0082] In a further variation, where a single broad wedge is used, with a
compound or other profile, it may be desirable to seat the wedge on two
or more springs in an inboard-and-outboard orientation to create a
restoring moment such as might not tend to be achieved by a single spring
alone. That is, even if a single large wedge is used, the use of two,
spaced apart springs may tend to generate a restoring moment if the wedge
tries to twist, since the deflection of one spring may then be greater
that the other.

[0083] When the dampers are placed in pairs, either immediately
side-by-side or with spacing between the pairs, the restoring moment for
squaring the truck will tend not only to be due to the increase in
compression to one set of springs due to the extra tendency to squeeze
the dampers downward in the pocket, but due to the difference in
compression between the springs that react to the extra squeezing of one
diagonal set of dampers and the springs that act against the opposite
diagonal pair that will tend to be less tightly squeezed.

SUMMARY OF THE INVENTION

[0084] In an aspect of the invention there is an autorack rail road car
having a car body for the transport of automobiles, the car body being
supported for rolling motion along rail road tracks by rail road car
trucks. At least one of the trucks has wheels whose diameter is greater
than 33 inches.

[0085] In a further feature of that aspect of the invention, at least one
of the trucks has wheels that are at least 36 inches in diameter. In
another feature of that aspect of the invention, the rail road car truck
has wheels that are at least 38 inches in diameter. In yet a further
feature of that aspect of the invention, at least one of the rail road
car trucks has an overall vertical spring rate of less than 50,000
Lbs./in. In a further feature, the overall vertical spring rate of the
truck is less than 40,000 Lbs./in. In a still further feature, the
overall vertical spring rate is less than 30,000 Lbs./in. In a still
further feature, the overall vertical spring rate is less than 20,000
Lbs./in. In a still further feature, the overall vertical spring rate is
in the range of 10,000 Lbs/in. to 20,000 Lbs./in.

[0086] In a still further feature, at least one of the trucks is a swing
motion truck. In an additional feature, the truck includes a pair of
first and second side frames and a transversely oriented truck bolster
mounted between the side frames. The side frames are mounted to the
wheelsets, and are able to swing laterally relative to the wheels. The
effective equivalent length of the swinging side frames is greater than
10 inches.

[0087] In a still further feature, at least one of the trucks is free of
unsprung lateral cross-members. In another feature of that feature of the
invention, the truck is free of a transom.

[0088] In still another feature of that aspect of the invention, at least
one of the trucks has friction dampers mounted in laterally spaced pairs,
the dampers being biased to exert a squaring restorative moment couple on
the truck bolster relative to the side frames when the truck bolster is
deflected from square relative to the side frames. In still another
feature of that aspect of the invention, at least one of the trucks has
springs mounted in inboard and outboard pairs between the bolster and
each of the side frames, said inboard and outboard pairs being oriented
to provide a squaring restorative moment couple to the bolster relative
to the side frames.

[0089] In still another feature of the invention, the rail car includes a
rail car body unit that has a weight of at least 90,000 Lbs., in an
unloaded condition. In a further feature of the invention, the rail car
body unit has an unladen weight of at least 100,000 Lbs. In another
further feature the rail car body unit has an unladen weight of at least
120,000 Lbs. In another further feature, the rail car body unit has an
unladen weight of at least 130,000 Lbs.

[0090] In another feature of that aspect of the invention, the rail road
car body unit includes at least 15,000 Lbs., of ballast. In another
feature, the rail road car body unit includes at least 25,000 Lbs., of
ballast. In another feature of the invention, the rail road car body unit
includes at least 40,000 Lbs., of ballast. In a further feature of the
invention, the ballast weight is incorporated in a deck plate. In another
feature of the invention the rail road car has a deck plate exceeding 3/8
inches in thickness. In another feature of the invention the rail road
car body has a deck plate exceeding 1/2 inches in thickness. In another
feature of the invention the rail road car body has a deck plate
exceeding 3/4 inches in thickness. In another feature of the invention
the rail road car body has a deck plate exceeding 1 inch in thickness. In
another feature of the invention the rail road car body has a deck plate
exceeding 11/4 inch in thickness.

[0091] In another feature of that aspect of the invention at least one of
the rail car trucks has a wheelbase exceeding 73 inches in length. In
another feature at least one of the trucks has a wheelbase that exceeds
1.3 times the gauge width of the rails. In another feature the wheelbase
is in the range of 78 to 88 inches in length. In another feature of the
invention the wheelbase is in the range of 1.3 to 1.6 times the track
gauge width.

[0092] In another feature of the invention, the rail road car is an
articulated railroad car. In still another feature of the invention, the
rail road car is an articulated rail road car, and one of the articulated
connectors is cantilevered relative to the truck closest thereto. In
another feature the articulated rail road car is a three pack rail road
car. In still another feature the three pack rail road car has a middle
unit connected between two end units. Each of the end units has a coupler
end truck, and each of the end units has an asymmetric car body weight
distribution in which most of the weight of the end car body is carried
by the end truck. In a further feature, the end car body is ballasted. In
a still further feature, the ballast of the end car body is has a
distribution that is biased toward the end truck.

BRIEF DESCRIPTION OF THE DRAWINGS

[0093] FIG. 1a shows a prior art exploded partial view illustration of a
swing motion truck, much as shown at page 716 in the 1980 Car and
Locomotive Cyclopedia;

[0094] FIG. 1b shows a cross-sectional detail of an upper rocker assembly
of the truck of FIG. 1a;

[0095] FIG. 1c shows a cross-sectional detail of a lower rocker assembly
of the truck of FIG. 1a;

[0096] FIG. 2a shows a side view of a single unit auto rack rail road car;

[0097] FIG. 2b shows a cross-sectional view of the auto-rack rail road car
of FIG. 2a in a bi-level configuration, one half section of FIG. 2b being
taken through the main bolster and the other half taken looking at the
cross-tie outboard of the main bolster;

[0098] FIG. 2c shows a half sectioned partial end view of the rail road
car of FIG. 2a illustrating the wheel clearance below the main deck, half
of the section being taken through the main bolster, the other half
section being taken outboard of the truck with the main bolster removed
for clarity;

[0099] FIG. 2d shows a partially sectioned side view of the rail road car
of FIG. 2c illustrating the relationship of the truck, the bolster and
the wheel clearance, below the main deck;

[0147] FIG. 17d shows an end view of the three piece truck of FIG. 17a.

[0148] FIG. 17e shows a schematic of a spring layout for the truck of FIG.
17a.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT

[0149] The description that follows, and the embodiments described
therein, are provided by way of illustration of an example, or examples,
of particular embodiments of the principles of the present invention.
These examples are provided for the purposes of explanation, and not of
limitation, of those principles and of the invention. In the description,
like parts are marked throughout the specification and the drawings with
the same respective reference numerals. The drawings are not necessarily
to scale and in some instances proportions may have been exaggerated in
order more clearly to depict certain features of the invention.

[0150] In terms of general orientation and directional nomenclature, for
each of the rail road cars described herein, the longitudinal direction
is defined as being coincident with the rolling direction of the car, or
car unit, when located on tangent (that is, straight) track. In the case
of a car having a center sill, whether a through center sill or stub
sill, the longitudinal direction is parallel to the center sill, and
parallel to the side sills, if any. Unless otherwise noted, vertical, or
upward and downward, are terms that use top of rail, TOR, as a datum. The
term lateral, or laterally outboard, refers to a distance or orientation
relative to the longitudinal centerline of the railroad car, or car unit,
indicated as CL-Rail Car. The term "longitudinally inboard", or
"longitudinally outboard" is a distance taken relative to a mid-span
lateral section of the car, or car unit. Pitching motion is angular
motion of a rail car unit about a horizontal axis perpendicular to the
longitudinal direction. Yawing is angular motion about a vertical axis.
Roll is angular motion about the longitudinal axis.

[0151] Reference is made in this description to rail car trucks and in
particular to three piece rail road freight car trucks. Several AAR
standard truck sizes are listed at page 711 in the 1997 Car & Locomotive
Cyclopedia. As indicated, for a single unit rail car having two trucks, a
"40 Ton" truck rating corresponds to a maximum gross car weight on rail
(GWR) of 142,000 lbs. Similarly, "50 Ton" corresponds to 177,000 lbs, "70
Ton" corresponds to 220,000 lbs, "100 Ton" corresponds to 263,000 lbs,
and "125 Ton" corresponds to 315,000 lbs. In each case the load limit per
truck is then half the maximum gross car weight on rail. Two other types
of truck are the "110 Ton" truck for 286,000 Lbs GWR and the "70 Ton
Special" low profile truck sometimes used for auto rack cars. Given that
the rail road car trucks described herein tend to have both longitudinal
and transverse axes of symmetry, a description of one half of an assembly
may generally also be intended to describe the other half as well,
allowing for differences between right hand and left hand parts.

[0152] Portions of this application refer to friction dampers, and
multiple friction damper systems. There are several types of damper
arrangement as shown at pages 715-716 of the 1997 Car and Locomotive
Encyclopedia, those pages being incorporated herein by reference. Double
damper arrangements are shown and described in my co-pending U.S. patent
application Ser. No. 10/210,797 now U.S. Pat. No. 6,895,866. Each of the
arrangements of dampers shown at pp. 715 to 716 of the 1997 Car and
Locomotive Encyclopedia can be modified to employ a four cornered, double
damper arrangement of inner and outer dampers.

[0153] FIGS. 2a, 3a, 3b, 4a, and 4b, show different types of rail road
freight cars in the nature of auto rack rail road cars, all sharing a
number of similar features. FIG. 2a (side view) shows a single unit
autorack rail road car, indicated generally as 20. It has a rail car body
22 supported for rolling motion in the longitudinal direction (i.e.,
along the rails) upon a pair of three-piece rail road freight car trucks
23 and 24 mounted at main bolsters at either of the first and second ends
26, 28 of rail car body 22. Body 22 has a housing structure 30, including
a pair of left and right hand sidewall structures 32, 34 and an
over-spanning canopy, or roof 36 that co-operate to define an enclosed
lading space. Body 22 has staging in the nature of a main deck 38 running
the length of the car between first and second ends 26, 28 upon which
wheeled vehicles, such as automobiles can be conducted by circus-loading.
Body 22 can have staging in either a bi-level configuration, as shown in
FIG. 2b, in which a second, or upper deck 40 is mounted above main deck
38 to permit two layers of vehicles to be carried; or a tri-level
configuration with a mid-level deck, similar to deck 40, and a top deck,
also similar to deck 40, are mounted above each other, and above main
deck 38 to permit three layers of vehicles to be carried. The staging,
whether bi-level or tri-level, is mounted to the sidewall structures 32,
34. Each of the decks defines a roadway, trackway, or pathway, by which
wheeled vehicles such as automobiles can be conducted between the ends of
rail road car 20.

[0154] A through center sill 50 extends between ends 26, 28. A set of
cross-bearers 52 extend to either side of center sill 50, terminating at
side sills 56, 58 that run the length of car 20 parallel to outer sill
50. Main deck 38 is supported above cross-bearers 52 and between side
sills 56, 58. Sidewall structures 32, 34 each include an array of
vertical support members, in the nature of posts 60, that extend between
side sills 56, 58, and top chords 62, 64. A corrugated sheet roof 66
extends between top chords 62 and 64 above deck 38 and such other decks
as employed. Radial arm doors 68, 70 enclose the end openings of the car,
and are movable to a closed position to inhibit access to the interior of
car 20, and to an open position to give access to the interior. Each of
the decks has bridge plate fittings (not shown) to permit bridge plates
to be positioned between car 20 and an adjacent car when doors 68 or 70
are opened to permit circus loading of the decks. Both ends of car 20
have couplers and draft gear for connecting to adjacent rail road cars.

Two-Unit Articulated Auto Rack Car

[0155] Similarly, FIG. 3a shows a two unit articulated auto rack rail road
car, indicated generally as 80. It has a first rail car unit body 82, and
a second rail car unit body 85, both supported for rolling motion in the
longitudinal direction (i.e., along the rails) upon rail car trucks 84,
86 and 88. Rail car trucks 84 and 88 are mounted at main bolsters at
respective coupler ends of the first and second rail car unit bodies 83
and 84. Truck 86 is mounted beneath articulated connector 90 by which
bodies 83 and 84 are joined together. Each of bodies 83 and 84 has a
housing structure 92, 93, including a pair of left and right hand
sidewall structures 94, 96 (or 95, 97) and a canopy, or roof 98 (or 99)
that define an enclosed lading space. A bellows structure 100 links
bodies 82 and 83 to discourage entry by vandals or thieves.

[0156] Each of bodies 82, 83 has staging in the nature of a main deck
similar to deck 38 running the length of the car unit between first and
second ends 104, 106 (105, 107) upon which wheeled vehicles, such as
automobiles can be conducted. Each of bodies 82, 83 can have staging in
either a bi-level configuration, as shown in FIG. 1b, or a tri-level
configuration. Other than brake fittings, and other minor fittings, car
unit bodies 82 and 83 are substantially the same, differing in that car
body 82 has a pair of female side-bearing arms adjacent to articulated
connector 90, and car body 83 has a co-operating pair of male side
bearing arms adjacent to articulated connector 90.

[0157] Each of car unit bodies 82 and 83 has a through center sill 110
that extends between the first and second ends 104, 106 (105, 107). A set
of cross-bearers 112, 114 extend to either side of center sill 110,
terminating at side sills 116, 118. Main deck 102 (or 103) is supported
above cross-bearers 112, 114 and between side sills 116, 118. Sidewall
structures 94, 96 and 95, 97 each include an array of vertical support
members, in the nature of posts 120, that extend between side sills 116,
118, and top chords 126, 128. A corrugated sheet roof 130 extends between
top chords 126 and 128 above deck 102 and such other decks as may be
employed.

[0158] Radial arm doors 132, 134 enclose the coupler end openings of car
bodies 82 and 83 of rail road car 80, and are movable to respective
closed positions to inhibit access to the interior of rail road car 80,
and to respective open positions to give access to the interior thereof.
Each of the decks has bridge plate fittings (upper deck fittings not
shown) to permit bridge plates to be positioned between car 80 and an
adjacent auto rack rail road car when doors 132 or 134 are opened to
permit circus loading of the decks.

[0159] For the purposes of this description, the cross-section of FIG. 2b
can be considered typical also of the general structure of the other
railcar unit bodies described below, whether 82, 85, 202, 204, 142, 144,
146, 222, 224 or 226. It should be noted that FIG. 2b shows a stepped
section in which the right hand portion shows the main bolster 75 and the
left hand section shows a section looking at the cross-tie 77 outboard of
the main bolster. The sections of FIGS. 2b and 2c are typical of the
sections of the end units described herein at their coupler end trucks,
such as trucks 232, 148, 84, 88, 210, 206. The upward recess in the main
bolster 75 provides vertical clearance for the side frames (typically 7''
or more). That is, the clearance `X` in FIG. 2c is about 7 inches in one
embodiment between the side frames and the bolster for an unladen car at
rest.

[0160] As may be noted, the web of main bolster 75 has a web rebate 79 and
a bottom flange that has an inner horizontal portion 69, an upwardly
stepped horizontal portion 71 and an outboard portion 73 that deepens to
a depth corresponding to the depth of the bottom flange of side sill 58.
Horizontal portion 69 is carried at a height corresponding generally to
the height of the bottom flange of side sill 58, and portion 71 is
stepped upwardly relative to the height of the bottom flange of side sill
58 to provide greater vertical clearance for the side frame of truck 23
or 24 as the case may be.

Three or More Unit Articulated Auto Rack Car

[0161] FIG. 4a shows a three unit articulated autorack rail road car,
generally as 140. It has a first end rail car unit body 142, a second end
rail car unit body 144, and an intermediate rail car unit body 146
between rail car unit bodies 142 and 144. Rail car unit bodies 142, 144
and 146 are supported for rolling motion in the longitudinal direction
(i.e., along the rails) upon rail car trucks 148, 150, 152, and 154. Rail
car trucks 148 and 150 are "coupler end" trucks mounted at main bolsters
at respective coupler ends of the first and second rail car bodies 142
and 144. Trucks 152 and 154 are "interior" or "intermediate" trucks
mounted beneath respective articulated connectors 156 and 158 by which
bodies 142 and 144 are joined to body 146. For the purposes of this
description, body 142 is the same as body 82, and body 144 is the same as
body 83. Rail car body 146 has a male end 159 for mating with the female
end 160 of body 142, and a female end 162 for mating with the male end
164 of rail car body 144.

[0162] Body 146 has a housing structure 166 like that of FIG. 2b, that
includes a pair of left and right hand sidewall structures 168 and a
canopy, or roof 170 that co-operate to define an enclosed lading space.
Bellows structures 172 and 174 link bodies 142, 146 and 144, 146
respectively to discourage entry by vandals or thieves.

[0163] Body 146 has staging in the nature of a main deck 176, similar to
deck 38, running the length of the car unit between first and second ends
178, 180 defining a roadway upon which wheeled vehicles, such as
automobiles can be conducted. Body 146 can have staging in either a
bi-level configuration or a tri-level configuration, to co-operate with
the staging of bodies 142 and 144.

[0164] Other than brake fittings, and other ancillary features, car bodies
142 and 144 are substantially the same, differing to the extent that car
body 142 has a pair of female side-bearing arms adjacent to articulated
connector 156, and car body 144 has a co-operating pair of male side
bearing arms adjacent to articulated connector 158.

[0165] Other articulated auto-rack cars of greater length can be assembled
by using a pair of end units, such as male and female end units 82 and
83, and any number of intermediate units, such as intermediate unit 146,
as may be suitable. In that sense, rail road car 140 is representative of
multi-unit articulated rail road cars generally.

Alternate Configurations

[0166] Alternate configurations of multi-unit rail road cars are shown in
FIGS. 3b and 4b. In FIG. 3b, a two unit articulated auto-rack rail road
car is indicated generally as 200. It has first and second rail car unit
bodies 202, 204 supported for rolling motion in the longitudinal
direction by three rail road car trucks, 206, 208 and 210 respectively.
Rail car unit bodies 202 and 204 are joined together at an articulated
connector 212. In this instance, while rail car bodies 202 and 204 share
the same basic structural features of rail car body 22, in terms of a
through center sill, cross-bearers, side sills, walls and canopy, and
vehicles decks, rail car body 202 is a "two-truck" body, and rail car
body 204 is a single truck body. That is, rail car body 202 has main
bolsters at both its first, coupler end, and at its second, articulated
connector end, the main bolsters being mounted over trucks 206 and 208
respectively. By contrast, rail car body 204 has only a single main
bolster, at its coupler end, mounted over truck 210. Articulated
connector 212 is mounted to the end of the respective center sills of
rail car bodies 202 and 204, longitudinally outboard of rail car truck
208. The use of a cantilevered articulation in this manner, in which the
pivot center of the articulated connector is offset from the nearest
truck center, is described more fully in my co-pending U.S. patent
application Ser. No. 09/614,815 for a Rail Road Car with Cantilevered
Articulation filed Jul. 12, 2000, incorporated herein by reference, now
U.S. Pat. No. 7,047,889, and may tend to permit a longer car body for a
given articulated rail road car truck center distance as therein
described.

[0167] FIG. 4b shows a three-unit articulated rail road car 220 having
first end unit 222, second end unit 224, and intermediate unit 226, with
cantilevered articulated connectors 228 and 230. End units 222 and 224
are single truck units of the same construction as car body 204.
Intermediate unit 226 is a two truck unit having similar construction to
car body 202, but having articulated connectors at both ends, rather than
having a coupler end. FIG. 4c shows an isometric view of end unit 224 (or
222). Analogous five pack articulated rail road cars having cantilevered
articulations can also be produced. Many alternate configurations of
multi-unit articulated rail road cars employing cantilevered
articulations can be assembled by re-arranging, or adding to, the units
illustrated.

[0168] In each of the foregoing descriptions, each of rail road cars 20,
80, 140, 200 and 220 has a pair of first and second coupler ends by which
the rail road car can be releasably coupled to other rail road cars,
whether those coupler ends are part of the same rail car body, or parts
of different rail car bodies of a multi-unit rail road car joined by
articulated connections, draw-bars, or a combination of articulated
connections and draw-bars.

[0169] FIGS. 5a and 5b show an example of a draft gear arrangement that
may be used at a first coupler end 300 of rail road car 20, coupler end
300 being representative of either of the coupler ends and draft gear
arrangement of rail road car 20, and of rail road cars 80, 140, 200 and
220 more generally. Coupler pocket 302 houses a coupler indicated as 304.
It is mounted to a coupler yoke 308, joined together by a pin 310. Yoke
308 houses a coupler follower 312, a draft gear 314 held in place by a
shim (or shims, as required) 316, a wedge 318 and a filler block 320.
Fore and aft draft gear stops 322, 324 are welded inside coupler pocket
302 to retain draft gear 314, and to transfer the longitudinal buff and
draft loads through draft gear 314 and on to coupler 304. In the
preferred embodiment, coupler 304 is an AAR Type F70DE coupler, used in
conjunction with an AAR Y45AE coupler yoke and an AAR Y47 pin. In the
preferred embodiment, draft gear 314 is a Mini-BuffGear such as
manufactured Miner Enterprises Inc., or by the Keystone Railway Equipment
Company, of 3420 Simpson Ferry Road, Camp Hill, Pa. As taken together,
this draft gear and coupler assembly yields a reduced slack, or low
slack, short travel, coupling as compared to an AAR Type E coupler with
standard draft gear or hydraulic EOCC device. As such it may tend to
reduce overall train slack. In addition to mounting the Mini-BuffGear
directly to the draft pocket, that is, coupler pocket 302, and hence to
the structure of the rail car body of rail road car 20, (or of the other
rail road cars noted above) the construction described and illustrated is
free of other long travel draft gear, sliding sills and EOCC devices, and
the fittings associated with them. The draft pocket arrangement may
include a flared bell-mouth and other features differing from the
illustrated example.

[0170] Mini-BuffGear has between 5/8 and 3/4 of an inch displacement
travel in buff at a compressive force greater than 700,000 Lbs. Other
types of draft gear can be used to give an official rating travel of less
than 21/2 inches under M-901-G, or if not rated, then a travel of less
than 2.5 inches under 500,000 Lbs. buff load. For example, while
Mini-BuffGear is preferred, other draft gear is available having a travel
of less than 13/4 inches at 400,000 Lbs., one known type has about 1.6
inches of travel at 400,000 Lbs., buff load. It is even more advantageous
for the travel to be less than 1.5 inches at 700,000 Lbs. buff load and,
as in the embodiment of FIGS. 5a and 5b, preferred that the travel be at
least as small as 1'' inches or less at 700,000 Lbs. buff load.

[0171] Similarly, while the AAR Type F70DE coupler is preferred, other
types of coupler having less than the 25/32'' (that is, less than about
3/4'') nominal slack of an AAR Type E coupler generally or the 20/32''
slack of an AAR E50ARE coupler can be used. In particular, in alternative
embodiments with appropriate housing changes where required, AAR Type
F79DE and Type F73BE (members of the Type F Family), with or without top
or bottom shelves; AAR Type CS; or AAR Type H couplers can be used to
obtain reduced slack relative to AAR Type E couplers.

[0172] In each of the examples herein, all of the trucks may have wheels
that are greater than 33 inches in diameter. The wheels can
advantageously be 36 inches or 38 inches in diameter, or possibly larger
depending on deck height geometry, and are preferred to be 36 inch
wheels. Although it is advantageous for the wheels of all of the trucks
to be of the same diameter, it is not necessary. That is, one or more
trucks, such as the intermediate truck or trucks in an articulated
autorack rail road car embodiment can have wheels of a larger diameter
than 33 inches such as 36 or 38 inches, for example, whereas the other
trucks, namely the end trucks can have 33 inch or other wheels.

Weight Distribution

[0173] In each of the autorack rail car embodiments described above, each
of the car units has a weight, that weight being carried by the rail car
trucks with which the car is equipped. In each of the embodiments of
articulated rail cars described above there is a number of rail car units
joined at a number of articulated connectors, and carried for rolling
motion along railcar tracks by a number of railcar trucks. In each case
the number of articulated car units is one more than the number of
articulations, and one less than the number of trucks. In the event that
some of the cars units are joined by draw bars the number of articulated
connections will be reduced by one for each draw bar added, and the
number of trucks will increase by one for each draw bar added. Typically
articulated rail road cars have only articulated connections between the
car units. All cars described have releasable couplers mounted at their
opposite ends.

[0174] In each case described above, where at least two car units are
joined by an articulated connector, there are end trucks (e.g. 150, 232)
inset from the coupler ends of the end car units, and intermediate trucks
(e.g. 154, 234) that are mounted closer to, or directly under, one or
other of the articulated connectors (e.g. 156, 230). In a car having
cantilevered articulations, such as shown in FIG. 36, the articulated
connector is mounted at a longitudinal offset distance (the cantilever
arm CA) from the truck center. In each case, each of the car units has an
empty weight, and also a full weight. The full weight is usually limited
by the truck capacity, whether 70 ton (33 inch diameter wheels), 100 ton
(36 inch diameter wheels), 110 ton (36 inch diameter wheels, 286,000
Lbs.) or 125 ton (38 inch diameter wheels). In some instances, with low
density lading, the volume of the lading is such that the truck loading
capacity cannot be reached without exceeding the volumetric capacity of
the car body.

[0175] The dead sprung weight of a rail car unit is generally taken as the
body weight of the car, including any ballast, as described below, plus
that portion of the weight of the truck bearing on the springs, that
portion most typically being the weight of the truck bolsters. The
unsprung weight of the trucks is, primarily, the weight of the side
frames, the axles and the wheels, plus ancillary items such as the
brakes, springs, and axle bearings. The unsprung weight of a three piece
truck may generally be about 8800 lbs. The live load is the weight of the
lading. The sum of (a) the live load; (b) the dead sprung load; and (c)
the unsprung weight of the trucks is the gross railcar weight on rail,
and is not to exceed the rated value for the truck.

[0176] In each of the embodiments described above, each of the rail car
units has a weight and a weight distribution of the dead sprung weight of
the car body which determines the dead sprung load carried by each truck.
In each of the embodiments described above, the sum of the sprung weights
of all of the car bodies of an articulated car is designated as WO.
(The sprung mass, MO, is the sprung weight WO divided by the
gravitational constant, g. In each case where a weight is given herein,
it is understood that conversion to mass can be readily made in this way,
particularly as when calculating natural frequencies). For a single unit,
symmetrical rail road car, such as car 20, the weight on both trucks is
equal. In all of the articulated auto rack rail road car embodiments
described above, the distributed sprung weight on any end truck, is at
least 2/3, and no more than 4/3 of the nearest adjacent interior truck,
such as an interior truck next closest to the nearest articulated
connector. It is advantageous that the dead sprung weight be in the range
of 4/5 to 6/5 of the dead sprung weight carried by the interior truck,
and it is preferred that the dead sprung weight be in the range of 90% to
110% of the interior truck. It is also desirable that the dead sprung
weight on any truck, WDS, fall in the range of 90% to 110% of the
value obtained by dividing WO by the total number of trucks of the
rail road car. Similarly, it is desirable that the dead sprung weight
plus the live load carried by each of the trucks be roughly similar such
that the overall truck loading is about the same. In any case, for the
embodiments described above, the design live load for one truck, such as
an end truck, can be taken as being at least 60% of the design live load
of the next adjacent truck, such as an internal truck. In terms of
overall dead and live loads, in each of the embodiments described the
overall sprung load of the end truck is at least 70% of the nearest
adjacent internal truck, advantageously 80% or more, and preferably 90%
of the nearest adjacent internal truck.

[0177] Inasmuch as the car weight would generally be more or less evenly
distributed on a lineal foot basis, and as such the interior trucks would
otherwise reach their load capacities before the coupler end trucks,
weight equalization may be achieved in the embodiments described above by
adding ballast to the end car units. That is, the dead sprung weight
distribution of the end car units is biased toward the coupler end, and
hence toward the coupler end truck (e.g. 84, 88, 206, 210, 150, 232). For
example, in the embodiments described above, a first ballast member is
provided in the nature of a main deck plate 350 of unusual thickness T
that forms part of main deck 38 of the rail car unit. Plate 350 extends
across the width of the end car unit, and from the longitudinally
outboard end of the deck a distance LB. In the embodiment of FIGS. 4b and
4c for example, the intermediate or interior truck 234 may be a 70 ton
truck near its sprung load limit of about 101,200 lbs., on the basis of
its share of loads from rail car units 222 and 226 (or, symmetrically 224
and 226 as the case may be), while, without ballast, end trucks 232 would
be at a significantly smaller sprung load, even when rail car 220 is
fully loaded. In this case, thickness T can be 11/2 inches, the width can
be 112 inches, and the length LB can be 312 inches, giving a weight of
roughly 15,220 lbs., centered on the truck center of end truck 232. This
gives a dead load of end car unit 222 of roughly 77,000 lbs., a dead
sprung load on end truck 232 of about 54,000 lbs., and a total sprung
load on truck 232 can be about 84,000 lbs. By comparison, center car unit
226 has a dead sprung load of about 60,000 lbs., with a dead sprung load
on interior truck 234 of about 55,000 lbs., and yielding a total sprung
load on interior truck 234 of 101,000 lbs when car 220 is fully loaded.
In this instance as much as a further 17,000 lbs. (±) of additional
ballast can be added before exceeding the "70 Ton" gross weight on rail
limit for the coupler end truck, 232. Ballast can also be added by
increasing the weight of the lower flange or webs of the center sill,
also advantageously reducing the center of gravity of the car. In
alternate embodiments plate thickness T can be a thickness greater than
1/2 inches, whether 3/4 inches, 1 inch, or 11/4 inches, or some other
thickness. Further, the ballast plate need not be a monolithic cut sheet,
but can be made up of a plurality of plates mounted at appropriate
locations to yield a mass (or weight) of ballast of suitable
distribution.

[0178] Similar weight distributions can be made for other capacities of
truck whether 100 Ton, 110 Ton or 125 Ton. With an increase in truck
capacity beyond "70 Ton", there is correspondingly an opportunity to add
more ballast up to the truck capacity limit. As noted above, although any
of these sizes of trucks can be used, it is preferable to use a truck
with a larger wheel diameter. That is, while 33 inch wheels (or even 28''
wheels in a "70 Ton Special") can be used, wheels larger than 33 inches
in diameter are preferred such as 36 inch or 38 inch wheels.

[0179] In the example of FIGS. 6a and 6b, a truck embodying an aspect of
the present invention is indicated as 410. Truck 410 differs from truck
A20 of FIG. 1a insofar as it is free of a rigid, unsprung lateral
connecting member in the nature of unsprung cross-bracing such as a frame
brace of crossed-diagonal rods, lateral rods, or a transom (such as
transom A60) running between the rocker plates of the bottom spring seats
of the opposed sideframes. Further, truck 410 employs gibs 412 to define
limits to the lateral range of travel of the truck bolster 414 relative
to the sideframe 416. In other respects, including the sideframe geometry
and upper and lower rocker assemblies, truck 410 is intended to have
generally similar features to truck A20, although it may differ in size,
pendulum length, spring stiffness, wheelbase, window width and window
height, and damping arrangement. The determination of these values and
dimensions may depend on the service conditions under which the truck is
to operate.

[0180] As with other trucks described herein, it will be understood that
since truck 410 (and trucks 420, 520, and 600, described below) are
symmetrical about both their longitudinal and transverse axes, the truck
is shown in partial section. In each case, where reference is made to a
sideframe, it will be understood that the truck has first and second
sideframes, first and second spring groups, and so on.

[0181] In FIGS. 7a and 7b, for example, a truck is identified generally as
420. Inasmuch as truck 420 is symmetrical about the truck center both
from side-to-side and lengthwise, the bolster, identified as 422, and the
sideframes, identified as 424 are shown in part. Truck 420 differs from
truck A20 of the prior art, described above, in that truck 420 has a
rigid bottom spring seat 444 rather than a lower rocker as in truck A20,
as described below, and is free of a rigid, unsprung lateral connection
member such as an underslung transom A60, a frame brace, or laterally
extending rods.

[0182] Sideframe 424 has a generally rectangular window 426 that
accommodates one of the ends 428 of the bolster 422. The upper boundary
of window 426 is defined by the sideframe arch, or compression member
identified as top chord member 430, and the bottom of window 426 is
defined by a tension member identified as bottom chord 432. The fore and
aft vertical sides of window 426 are defined by sideframe columns 434.

[0183] The ends of the tension member sweep up to meet the compression
member. At each of the swept-up ends of sideframe 424 there are sideframe
pedestal fittings 438. Each fitting 438 accommodates an upper rocker
identified as a pedestal rocker seat 440. Pedestal rocker seat 440
engages the upper surface of a bearing adapter 442. Bearing adapter 442
engages a bearing mounted on one of the axles of the truck adjacent one
of the wheels. A rocker seat 440 is located in each of the fore and aft
pedestal fittings 438, the rocker seats 440 being longitudinally aligned
such that the sideframe can swing transversely relative to the rolling
direction of the truck in a "swing hanger" arrangement.

[0184] Bearing adapter 442 has a hollowed out recess 441 in its upper
surface that defines a bearing surface for receiving rocker seat 440.
Bearing surface 441 is formed on a radius of curvature R1. The
radius of curvature R1 is preferably in the range of less than 25
inches, may be in the range of 5'' to 15'', and is preferably in the
range of 8 to 12 inches, and most preferably about 10 inches with the
center of curvature lying upwardly of the rocker seat. The lower face of
rocker seat 440 is also formed on a circular arc, having a radius of
curvature R2 that is less than the radius of curvature R1 of
the recess of surface recess 441. R2 is preferably in the range of
1/4 to 3/4 as large as R1, and is preferably in the range of 3-10
inches, and most preferably 5 inches when R1 is 10 inches, i.e.,
R2 is one half of R1. Given the relatively small angular
displacement of the rocking motion of R2 relative to R1
(typically less than ±10 degrees) the relationship is one of rolling
contact, rather than sliding contact.

[0185] The bottom chord or tension member of sideframe 424 has a basket
plate, or lower spring seat 444 rigidly mounted to bottom chord 432, such
that it has a rigid orientation relative to window 426, and to sideframe
424 in general. That is, in contrast to the lower rocker platform of the
prior art swing motion truck A20 of FIG. 1a, as described above, spring
seat 444 is not mounted on a rocker, and does not rock relative to
sideframe 424. Although spring seat 444 retains an array of bosses 446
for engaging the corner elements 454, namely springs 454 and 455
(inboard), 456 and 457 (outboard) of a spring set 448, there is no
transom mounted between the bottom of the springs and seat 444. Seat 444
has a peripheral lip 452 for discouraging the escape of the bottom ends
the of springs.

[0186] The spring group, or spring set 448, is captured between the distal
end 428 of bolster 422 and spring seat 444, being placed under
compression by the weight of the rail car body and lading that bears upon
bolster 422 from above.

[0187] Friction damping is provided by damping wedges 462 that seat in
mating bolster pockets 464 that have inclined damper seats 466. The
vertical sliding faces 470 of the friction damper wedges 462 then ride up
and down on friction wear plates 472 mounted to the inwardly facing
surfaces of sideframe columns 434. Angled faces 474 of wedges 462 ride
against the angled face of seat 466. Bolster 422 has inboard and outboard
gibs 476, 478 respectively, that bound the lateral motion of bolster 422
relative to sideframe columns 434. This motion allowance may
advantageously be in the range of ±11/8 to 13/4 inches, and is most
preferably in the range of 1 3/16 to 1 9/16 inches, and can be set, for
example, at 11/2 inches or 11/4 inches of lateral travel to either side
of a neutral, or centered, position when the sideframe is undeflected.

[0188] As in the prior art swing motion truck A20, a spring group of 8
springs in a 3:2:3 arrangement is used. Other configurations of spring
groups could be used, such as those described below.

[0189] In the embodiment of FIG. 8, a truck 520 is substantially similar
to truck 420, but differs insofar as truck 520 has a bolster 522 having
double bolster pockets 524, 526 on each face of the bolster at the
outboard end. Bolster pockets 524, 526 accommodate a pair of first and
second, laterally inboard and laterally outboard friction damper wedges
528, 529 and 530, 531, respectively. Wedges 528, 529 each sit over a
first, inboard corner spring 532, 533, and wedges 530, 531 each sit over
a second, outboard corner spring 534, 535. In this four corner
arrangement, each damper is individually sprung by one or another of the
springs in the spring group. The static compression of the springs under
the weight of the car body and lading tends to act as a spring loading to
bias the damper to act along the slope of the bolster pocket to force the
friction surface against the sideframe. As such, the dampers co-operate
in acting as biased members working between the bolster and the side
frames to resist parallelogram, or lozenging, deformation of the side
frame relative to the truck bolster. A middle end spring 536 bears on the
underside of a land 538 located intermediate bolster pockets 524 and 526.
The top ends of the central row of springs, 540, seat under the main
central portion 542 of the end of bolster 522.

[0190] The lower ends of the springs of the entire spring group,
identified generally as 544, seat in the lower spring seat 546. Lower
spring seat 546 has the layout of a tray with an upturned rectangular
peripheral lip. Lower spring seat 546 is rigidly mounted to the lower
chord 548 of sideframe 549. In this case, spring group 544 has a 3
rows×3 columns layout, rather than the 3:2:3 arrangement of truck
420. A 3×5 layout as shown in FIG. 17e (described below) could be
used, as could other alternate spring group layouts. Truck 520 is free of
any rigid, unsprung lateral sideframe connection members such as transom
A60.

[0191] It will be noted that bearing plate 550 mounted to vertical
sideframe columns 552 is significantly wider than the corresponding
bearing plate 472 of truck 420 of FIG. 6a. This additional width
corresponds to the additional overall damper span width measured fully
across the damper pairs, plus lateral travel as noted above, typically
allowing roughly 11/2 (±) inches of lateral travel (i.e. for an
overall total of roughly 3'' travel) of the bolster relative to the
sideframe to either side of the undeflected central position. That is,
rather than having the width of one coil, plus allowance for travel,
plate 550 has the width of three coils, plus allowance to accommodate
11/2 (±) inches of travel to either side. Plate 550 is significantly
wider than the through thickness of the sideframes more generally, as
measured, for example, at the pedestals.

[0192] Damper wedges 528 and 530 sit over 44% (±) of the spring group
i.e., 4/9 of a 3 rows×3 columns group as shown in FIG. 8, whereas
wedges 470 only sat over 2/8 of the 3:2:3 group in FIG. 7a. For the same
proportion of vertical damping, wedges 528 and 530 may tend to have a
larger included angle (i.e., between the wedge hypotenuse and the
vertical face for engaging the friction wear plates on the sideframe
columns 434. For example, if the included angle of friction wedges 472 is
about 35 degrees, then, assuming a similar overall spring group
stiffness, and single coils, the corresponding angle of wedges 528 and
530 could advantageously be in the range of 50-65 degrees, or more
preferably about 55 degrees. In a 3×5 group such as group 976 of
truck 970 of FIG. 17e, for coils of equal stiffness, the wedge angle may
tend to be in the 35 to 40 degree range. The specific angle will be a
function of the specific spring stiffnesses and spring combinations
actually employed.

[0193] The use of spaced apart pairs of dampers 528, 530 may tend to give
a larger moment arm, as indicated by dimension "2M", for resisting
parallelogram deformation of truck 520 more generally as compared to
trucks 420 or A20. Parallelogram deformation may tend to occur, for
example, during the "truck hunting" phenomenon that has a tendency to
occur in higher speed operation.

[0194] Placement of doubled dampers in this way may tend to yield a
greater restorative "squaring" force to return the truck to a square
orientation than for a single damper alone, as in truck 420. That is, in
parallelogram deformation, or lozenging, the differential compression of
one diagonal pair of springs (e.g., inboard spring 532 and outboard
spring 535 may be more pronouncedly compressed) relative to the other
diagonal pair of springs (e.g., inboard spring 533 and outboard spring
534 may be less pronouncedly compressed than springs 532 and 535) tends
to yield a restorative moment couple acting on the sideframe wear plates.
This moment couple tends to rotate the sideframe in a direction to square
the truck, (that is, in a position in which the bolster is perpendicular,
or "square", to the sideframes) and thus may tend to discourage the
lozenging or parallelogramming, noted by Weber.

[0195] FIGS. 9a, 9b, 9c, 9d and 9e all relate to a three piece truck 600
for use with the rail road cars of FIG. 2a, 3a, 3b, 4a or 4b. FIGS. 2c
and 2d show the relationship of this truck to the deck level of these
rail road cars. Truck 600 has three major elements, those elements being
a truck bolster 602, symmetrical about the truck longitudinal centerline,
and a pair of first and second side frames, indicated as 604. Only one
side frame is shown in FIG. 9c given the symmetry of truck 600. Three
piece truck 600 has a resilient suspension (a primary suspension)
provided by a spring groups 605 trapped between each of the distal (i.e.,
transversely outboard) ends of truck bolster 602 and side frames 604.

[0196] Truck bolster 602 is a rigid, fabricated beam having a first end
for engaging one side frame assembly and a second end for engaging the
other side frame assembly (both ends being indicated as 606). A center
plate or center bowl 608 is located at the truck center. An upper flange
610 extends between the two ends 606, being narrow at a central waist and
flaring to a wider transversely outboard termination at ends 606. Truck
bolster 602 also has a lower flange 612 and two fabricated webs 614
extending between upper flange 610 and lower flange 612 to form an
irregular, closed section box beam. Additional webs 615 are mounted
between the distal portions of upper flange 610 and 614 where bolster 602
engages one of the spring groups 605. The transversely distal region of
truck bolster 602 also has friction damper seats 616, 618 for
accommodating friction damper wedges as described further below.

[0197] Side frame 604 is a casting having bearing seats 619 into which
bearing adapters 620, bearings 621, and a pair of axles 622 mount. Each
of axles 622 has a pair of first and second wheels 623, 625 mounted to it
in a spaced apart position corresponding to the width of the track gauge
of the track upon which the rail car is to operate. Side frame 604 also
has a compression member, or upper beam member 624, a tension member, or
lower beam member 626, and vertical side columns 628 and 630, each lying
to one side of a vertical transverse plane 625 bisecting truck 600 at the
longitudinal station of the truck center. A generally rectangular opening
in the nature of a sideframe window 627 is defined by the co-operation of
the upper and lower beam members 624, 626 and vertical columns 628, 630.
The distal end of truck bolster 602 can be introduced into window 627.
The distal end of truck bolster 602 can then move up and down relative to
the side frame within this opening. Lower beam member 626 (the tension
member) has a bottom or lower spring seat 632 upon which spring group 605
can seat. Similarly, an upper spring seat 634 is provided by the
underside of the distal portion of bolster 602 to engages the upper end
of spring group 605. As such, vertical movement of truck bolster 602 will
tend to compress or release the springs in spring group 605.

[0198] For the purposes of this description the swiveling, 4 wheel, 2 axle
truck 600 has first and second sideframes 604 that can be taken as having
the same upper rocker assembly as truck 520, and has a rigidly mounted
lower spring seat 632, like spring seat 544, but having a shape to suit
the 2 rows×4 columns spring layout rather than the 3×3 layout
of truck 520. It may also be noted that sideframe window 627 has greater
width between sideframe columns 628, 630 than window 526 between columns
528 to accommodate the longer spring group footprint, and bolster 602
similarly has a wider end to sit over the spring group.

[0199] In the embodiment of FIG. 9a, spring group 605 has two rows of
springs 636, a transversely inboard row and a transversely outboard row,
each row having four large (8 inch ±) diameter coil springs 636, 637,
638, 639 giving vertical bounce spring rate constant, k, for group 605 of
less than 10,000 lbs/inch. This spring rate constant can be in the range
of 6000 to 10,000 lbs/in., and is advantageously in the range of 7000 to
9500 lbs/in, and preferably in the range of 8000-8500 lbs./in., giving an
overall vertical bounce spring rate for the truck of double these values,
preferably in the range of 14000 to 18,500 lbs/in, or more narrowly,
16,000-17000 lbs./in. for the truck. The spring array can include nested
coils of outer springs, inner springs, and inner-inner springs depending
on the overall spring rate desired for the group, and the apportionment
of that stiffness. The number of springs, the number of inner and outer
coils, and the spring rate of the various springs can be varied. The
spring rates of the coils of the spring group add to give the spring rate
constant of the group, typically being suited for the loading for which
the truck is designed.

[0200] Each side frame assembly also has four friction damper wedges
arranged in first and second pairs of transversely inboard and
transversely outboard wedges 640, 641, 642 and 643 that engage the
sockets, or seats 616, 618 in a four-cornered arrangement. The corner
springs in spring group 605 bear upon a friction damper wedge 640, 641,
642 or 643. Each of vertical columns 628, 630 has a friction wear plate
650 having transversely inboard and transversely outboard regions against
which the friction faces of wedges 640, 641, 642 and 643 can bear,
respectively. Bolster gibs 651 and 653 lie inboard and outboard of wear
plate 650 respectively. Gibs 651 and 653 act to limit the lateral travel
of bolster 602 relative to side frame 604. The deadweight compression of
the springs under the dampers will tend to yield a reaction force working
on the bottom face of the wedge, trying to drive the wedge upward along
the inclined face of the seat in the bolster, thus urging, or biasing,
the friction face against the opposing portion of the friction face of
the side frame column. In one embodiment, the springs chosen can have an
undeflected length of 15 inches, and a dead weight deflection of about 3
inches.

[0201] As seen in the top view of FIG. 9c, and in the schematic sketch of
FIG. 9f the side-by-side friction dampers have a relatively wide averaged
moment arm L to resist angular deflection of the side frame relative to
the truck bolster in the parallelogram mode. This moment arm is
significantly greater than the effective moment arm of a single wedge
located on the spring group (and side frame) centre line. Further, the
use of independent springs under each of the wedges means that whichever
wedge is jammed in tightly, there is always a dedicated spring under that
specific wedge to resist the deflection. In contrast to older designs,
the overall damping face width is greater because it is sized to be
driven by relatively larger diameter (e.g., 8 in ±) springs, as
compared to the smaller diameter of, for example, AAR B 432 out or B 331
side springs, or smaller. Further, in having two elements side-by-side
the effective width of the damper is doubled, and the effective moment
arm over which the diagonally opposite dampers work to resist
parallelogram deformation of the truck in hunting and curving greater
than it would have been for a single damper.

[0202] In the illustration of FIG. 9e, the damper seats are shown as being
segregated by a partition 652. If a longitudinal vertical plane 654 is
drawn through truck 600 through the center of partition 652, it can be
seen that the inboard dampers lie to one side of plane 654, and the
outboard dampers lie to the outboard side of plane 654. In hunting then,
the normal force from the damper working against the hunting will tend to
act in a couple in which the force on the friction bearing surface of the
inboard pad will always be fully inboard of plane 654 on one end, and
fully outboard on the other diagonal friction face. For the purposes of
conceptual visualization, the normal force on the friction face of any of
the dampers can be idealized as an evenly distributed pressure field
whose effect can be approximated by a point load whose magnitude is equal
to the integrated value of the pressure field over its area, and that
acts at the centroid of the pressure field. The center of this
distributed force, acting on the inboard friction face of wedge 640
against column 628 can be thought of as a point load offset transversely
relative to the diagonally outboard friction face of wedge 643 against
column 630 by a distance that is notionally twice dimension `L` shown in
the conceptual sketch of FIG. 9f. In the example, this distance is about
one full diameter of the large spring coils in the spring set. It is a
significantly greater effective moment arm distance than found in typical
friction damper wedge arrangements. The restoring moment in such a case
would be, conceptually, MR=[(F1+F3)-(F2+F4)]L.
As indicated by the formulae on the conceptual sketch of FIG. 9f, the
difference between the inboard and outboard forces on each side of the
bolster is proportional to the angle of deflection ε of the truck
bolster relative to the side frame, and since the normal forces due to
static deflection x0 may tend to cancel out, MR=4kc
Tan(ε)Tan(θ)L, where θ is the primary angle of the
damper, and kc is the vertical spring constant of the coil upon
which the damper sits and is biased.

[0203] Further, in typical friction damper wedges, the enclosed angle of
the wedge tends to be somewhat less than 35 degrees measured from the
vertical face to the sloped face against the bolster. As the wedge angle
decreases toward 30 degrees, the tendency of the wedge to jam in place
increases. Conventionally the wedge is driven by a single spring in a
large group. The portion of the vertical spring force acting on the
damper wedges can be less than 15% of the group total. In the embodiment
of FIG. 9b, it is 50% of the group total (i.e., 4 of 8 equal springs).
The wedge angle of wedges 640, 642 is significantly greater than 35
degrees. The use of more springs, or more precisely a greater portion of
the overall spring stiffness, under the dampers, permits the enclosed
angle of the wedge to be over 35 degrees, whether in the range of between
roughly 37 to 40 or 45 degrees, to roughly 60 or 65 degrees.

[0204] In this example, damper wedges 640, 641 and 642, 643 sit over 50%
of the spring group i.e., 4/8 namely springs 636, 637, 638, 639. For the
same proportion of vertical damping as in truck 420, wedges 640, 641 and
642, 643 may tend to have a larger included angle, possibly about 60
degrees, although angles in the range of 45 to 70 degrees could be chosen
depending on spring combinations and spring stiffnesses. Once again, in a
warping condition, the somewhat wider damping region (the width of two
full coils plus lateral travel of 11/2'' (+/-)) of sideframe column wear
plates 627, 629 lying between inboard and outboard gibs 611, 613, 615,
617 relative to truck 20 (a damper width of one coil with travel), sprung
on individual springs (inboard and outboard in truck 600, as opposed to a
single central coil in truck 20), may tend to generate a moment couple to
give a restoring force working on a moment arm. This restoring force may
tend to urge the sideframe back to a square orientation relative to the
bolster, with diagonally opposite pairs of springs working as described
above. In this instance, the springs each work on a moment arm distance
corresponding to half of the distance between the centers of the 2 rows
of coils, rather than half the 3 coil distance shown in FIG. 8.

[0205] Where a softer suspension is used employing a relatively small
number of large diameter springs, such as in a 2×4, 3×3, or
3×5 group as described in the detailed description of the invention
herein, dampers may be mounted over each of four corner positions. In
that case, the portion of spring force acting under the damper wedges may
be in the 25-50% range for springs of equal stiffness. If the coils or
coil groups are not of equal stiffness, the portion of spring force
acting under the dampers may be in the range of perhaps 20% to 70%. The
coil groups can be of unequal stiffness if inner coils are used in some
springs and not in others, or if springs of differing spring constant are
used.

[0206] The size of the spring group embodiment of FIG. 9b yields a side
frame window opening having a width between the vertical columns of side
frame 604 of roughly 33 inches. This is relatively large compared to
existing spring groups, being more than 25% greater in width. In an
alternate 3×5 spring group arrangement of 51/2'' diameter springs,
the opening between the sideframe columns is more than 271/2 inches wide,
in one preferred embodiment being between 29 and 30 inches wide, namely
about 291/4 inches.

[0207] Truck 600 has a correspondingly greater wheelbase length, indicated
as WB. WB is advantageously greater than 73 inches, or, taken as a ratio
to the track gauge width, is advantageously greater than 1.30 time the
track gauge width. It is preferably greater than 80 inches, or more than
1.4 times the gauge width, and in one embodiment is greater than 1.5
times the track gauge width, being as great, or greater than, about 86
inches. Similarly, the side frame window is advantageously wider than
tall, the measurement across the wear plate faces of the side frame
columns being advantageously greater than 24'', possibly in the ratio of
greater than 8:7 of width to height, and possibly in the range of 28'' or
32'' or more, giving ratios of greater than 4:3 and greater than 3:2. The
spring seat may have lengthened dimensions to correspond to the width of
the side frame window, and a transverse width of 151/2''-17'' or more.

[0208] In FIGS. 10a, 10b and 10c, there is an alternate embodiment of soft
spring rate, long wheelbase three piece truck, identified as 660. Truck
660 employs constant force inboard and outboard, fore and aft pairs of
friction dampers 666 mounted in the distal ends of truck bolster 668. In
this arrangement, springs 670 are mounted horizontally in pockets in the
distal ends of truck bolster 668 and urge, or bias, each of the friction
dampers 666 against the corresponding friction surfaces of the vertical
columns of the side frames.

[0209] The spring force on friction damper wedges 640, 641, 642 and 643
varies as a function of the vertical displacement of truck bolster 602,
since they are driven by the vertical springs of spring group 605. By
contrast, the deflection of springs 670 does not depend on vertical
compression of the main spring group 672, but rather is a function of an
initial pre-load. Although the arrangement of FIGS. 10a, 10b and 10c
still provides inboard and outboard dampers and independent springing of
the dampers, the embodiment of FIG. 9b is preferred to that of FIGS. 6a,
6b and 6c.

Damper Variations

[0210] FIGS. 11a and 11b show a partial isometric view of a truck bolster
680 that is generally similar to truck bolster 600 of FIG. 9a, except
insofar as bolster pocket 682 does not have a central partition like web
652, but rather has a continuous bay extending across the width of the
underlying spring group, such as spring group 636. A single wide damper
wedge is indicated as 684. Damper 684 is of a width to be supported by,
and to be acted upon, by two springs 686, 688 of the underlying spring
group. In the event that bolster 600 may tend to deflect to a
non-perpendicular orientation relative to the associated side frame, as
in the parallelogramming phenomenon, one side of wedge 684 will tend to
be squeezed more tightly than the other, giving wedge 684 a tendency to
twist in the pocket about an axis of rotation perpendicular to the angled
face (i.e., the hypotenuse face) of the wedge. This twisting tendency may
also tend to cause differential compression in springs 686, 688, yielding
a restoring moment both to the twisting of wedge 684 and to the
non-square displacement of truck bolster 680 relative to the truck side
frame. As there may tend to be a similar moment generated at the opposite
spring pair at the opposite side column of the side frame, this may tend
to enhance the self-squaring tendency of the truck more generally.

[0211] Also included in FIG. 11b is an alternate pair of damper wedges
690, 692. This dual wedge configuration can similarly seat in bolster
pocket 682, and, in this case, each wedge 690, 692 sits over a separate
spring. Wedges 690, 692 are in a side-by-side independently displaceable
vertically slidable relationship relative to each other along the primary
angle of the face of bolster pocket 682. When the truck moves to an out
of square condition, differential displacement of wedges 690, 692 may
tend to result in differential compression of their associated springs,
e.g., 686, 688 resulting in a restoring moment as above.

[0212] The sliding motion described above may tend to cause wear on the
moving surfaces, namely (a) the side frame columns, and (b) the angled
surfaces of the bolster pockets. To alleviate, or ameliorate, this
situation, consumable wear plates 694 can be mounted in bolster pocket
682 (with appropriate dimensional adjustments) as in FIG. 11b. Wear
plates 694 can be smooth steel plates, possibly of a hardened, wear
resistant alloy, or can be made from a non-metallic, or partially
non-metallic, relatively low friction wear resistant surface. Other
plates for engaging the friction surfaces of the dampers can be mounted
to the side frame columns, and indicated by item 696 in FIG. 16a.

[0213] For the purposes of this example, it has been assumed that the
spring group is two coils wide, and that the pocket is, correspondingly,
also two coils wide. The spring group could be more than two coils wide.
The bolster pocket is assumed to have the same width as the spring group,
but could be less wide. For two coils where in some embodiments the group
may be more than two coils wide. A symmetrical arrangement of the dampers
relative to the side frame and the spring group is desirable, but an
asymmetric arrangement could be made. In the embodiments of FIGS. 9a, 11a
and 17a, the dampers are in four cornered arrangements that are
symmetrical both about the center axis of the truck bolster and about a
longitudinal vertical plane of the side frame.

[0214] Similarly, the wedges themselves can be made from a relatively
common material, such as a mild steel, and the given consumable wear face
members in the nature of shoes, or wear members. Such an arrangement is
shown in FIG. 12 in which a damper wedge is shown generically as 700. The
replaceable, consumable wear members are indicated as 702, 704. The
wedges and wear members have mating male and female mechanical interlink
features, such as the cross-shaped relief 703 formed in the primary
angled and vertical faces of wedge 700 for mating with the corresponding
raised cross shaped features 705 of wear members 702, 704. Sliding wear
member 702 is preferably made of a non-metallic, low friction material.

[0215] Although FIG. 12 shows a consumable insert in the nature of a wear
plate, the entire bolster pocket can be made as a replaceable part, as in
FIG. 11a. This bolster pocket can be made of a high precision casting, or
can be a sintered powder metal assembly having desired physical
properties. The part so formed is then welded into place in the end of
the bolster, as at 706 indicated in FIG. 11a.

[0216] The underside of the wedges described herein, wedge 700 being
typical in this regard, has a seat, or socket 707, for engaging the top
end of the spring coil, whichever spring it may be, spring 762 being
shown as typically representative. Socket 707 serves to discourage the
top end of the spring from wandering away from the intended generally
central position under the wedge. A bottom seat, or boss for discouraging
lateral wandering of the bottom end of the spring is shown in FIG. 16a as
item 708.

[0217] Thus far only primary angles have been discussed. FIG. 11c shows an
isometric view of an end portion of a truck bolster 710, generally
similar to bolster 600. As with all of the truck bolsters shown and
discussed herein, bolster 710 is symmetrical about the longitudinal
vertical plane of the bolster (i.e., cross-wise relative to the truck
generally) and symmetrical about the vertical mid-span section of the
bolster (i.e., the longitudinal plane of symmetry of the truck generally,
coinciding with the rail car longitudinal center line). Bolster 710 has a
pair of spaced apart bolster pockets 712, 714 for receiving damper wedges
716, 718. Pocket 712 is laterally inboard of pocket 714 relative to the
side frame of the truck more generally. Consumable wear plate inserts
720, 722 are mounted in pockets 712, 714 along the angled wedge face.

[0218] As can be seen, wedges 716, 718 have a primary angle, α as
measured between vertical sliding face 724, (or 726, as may be) and the
angled vertex 728 of outboard face 730. For the embodiments discussed
herein, primary angle α will tend to be greater than 40 degrees,
and may typically lie in the range of 45-65 degrees, possibly about 55-60
degrees. This angle will be common to the slope of all points on the
sliding hypotenuse face of wedge 716 (or 718) when taken in any plane
parallel to the plane of outboard end face 730. This same angle α
is matched by the facing surface of the bolster pocket, be it 712 or 714,
and it defines the angle upon which displacement of wedge 716, (or 718)
is intended to move relative to that surface.

[0219] A secondary angle β gives the inboard, (or outboard), rake of
the hypotenuse surface of wedge 716 (or 718). The true rake angle can be
seen by sighting along plane of the hypotenuse face and measuring the
angle between the hypotenuse face and the planar outboard face 730. The
rake angle is the complement of the angle so measured. The rake angle may
tend to be greater than 5 degrees, may lie in the range of 10 to 20
degrees, and is preferably about 15 degrees. A modest angle is desirable.

[0220] When the truck suspension works in response to track perturbations,
the damper wedges may tend to work in their pockets. The rake angles
yield a component of force tending to bias the outboard face 730 of
outboard wedge 718 outboard against the opposing outboard face of bolster
pocket 714. Similarly, the inboard face of wedge 716 will tend to be
biased toward the inboard planar face of inboard bolster pocket 712.
These inboard and outboard faces of the bolster pockets are preferably
lined with a low friction surface pad, indicated generally as 732. The
left hand and right hand biases of the wedges may tend to keep them apart
to yield the full moment arm distance intended, and, by keeping them
against the planar facing walls, may tend to discourage twisting of the
dampers in the respective pockets.

[0221] Bolster 710 includes a middle land 734 between pockets 712, 714,
against which another spring 736 may work, such as might be found in a
spring group that is three (or more) coils wide. However, whether two,
three, or more coils wide, and whether employing a central land or no
central land, bolster pockets can have both primary and secondary angles
as illustrated in the example embodiment of FIG. 11c, with or without
(though preferably with) wear inserts.

[0222] In the case where a central land, such as land 734 separates two
damper pockets, the opposing wear plates of the side frame columns need
not be monolithic. That is, two wear plate regions could be provided, one
opposite each of the inboard and outboard dampers, presenting planar
surfaces against which those dampers can bear. Advantageously, the normal
vectors of those regions are parallel, and most conveniently those
surfaces are co-planar and perpendicular to the long axis of the side
frame, and present a clear, un-interrupted surface to the friction faces
of the dampers.

[0223] The examples of FIGS. 11a, 11b and 11c are arranged in order of
incremental increases in complexity. The Example of FIG. 11d again
provides a further incremental increase in complexity. FIG. 11d shows a
bolster 740 that is similar to bolster 710 except insofar as bolster
pockets 742, 744 each accommodate a pair of split wedges 746, 748.
Pockets 742, 744 each have a pair of bearing surfaces 750, 752 that are
inclined at both a primary angle and a secondary angle, the secondary
angles of surfaces 750 and 752 being of opposite hand to yield the damper
separating forces discussed above. Surfaces 750 and 752 are also provided
with linings in the nature of relatively low friction wear plates 754,
756. Each of pockets 742 and 744 accommodates a pair of split wedges 758,
760. Each pair of split wedges seats over a single spring 762. Another
spring 764 bears against central land 766.

[0224] The example of FIG. 13a shows a combination of a bolster 770 and
biased split wedges 772, 774. Bolster 770 is the same as bolster 740
except insofar as bolster pockets 776, 778 are stepped pockets in which
the steps, e.g., items 780, 782, have the same primary angle, and the
same secondary angle, and are both biased in the same direction, unlike
the symmetrical sliding faces of the split wedges in FIG. 11d, which are
left and right handed. Thus the outboard pair of split wedges 784 has a
first member 786 and a second member 788 each having primary angle
α and secondary angle β, and are of the same hand such that in
use both the first and second members will tend to be biased in the
outboard direction (i.e. toward the distal end of bolster 770).
Similarly, the inboard pair of split wedges 790 has a first member 792
and a second member 794 each having primary angle α, and secondary
angle β, except that the sense of secondary angle β is in the
opposite direction such that members 792 and 792 will tend in use to be
driven in the inboard direction (i.e., toward the truck center).

[0225] As shown in the partial sectional view of FIG. 13c, a replaceable
monolithic stepped wear insert 796 is welded in the bolster pocket 780
(or 782 if opposite hand, as the case may be). Insert 796 has the same
primary and secondary angles α and β as the split wedges it is
to accommodate, namely 786, 788 (or, opposite hand, 792, 794). When
installed, and working, the more outboard of the wedges, 788 (or,
opposite hand, the more inboard of the wedges 792) has a vertical and
longitudinally planar outboard face 800 that bears against a similarly
planar outboard face 802 (or, opposite hand, inboard face 804) These
faces are preferably prepared in a manner that yields a relatively low
friction sliding interface between them. In that regard, a low friction
pad may be mounted to either surface, preferably the outboard surface of
pocket 780. The hypotenuse face 806 of member 788 bears against the
opposing outboard land 810 of insert 796. The overall width of outboard
member 788 is greater than that of outboard land 810, such that the
inboard planar face of member 788 acts as an abutment face to fend
inboard member 786 off of the surface of the step 812 in insert 796.

[0226] In similar manner inboard wedge member 786 has a hypotenuse face
814 that bears against the inboard land portion 816 of insert 796. The
total width of bolster pocket 780 is greater than the combined width of
wedge members, such that a gap is provided between the inboard
(non-contacting) face of member 786 and the inboard planar face of pocket
780. The same relationship, but of opposite hand, exists between pocket
782 and members 792, 794.

[0227] In an optional embodiment, a low friction pad, or surfacing, can be
used at the interface of members 786, 788 (or 792, 794) to facilitate
sliding motion of the one relative to the other.

[0228] In this arrangement, working of the wedges, i.e., members 786, 788
against the face of insert 796 will tend to cause both members to move in
one direction, namely to their most outboard position. Similarly, members
792 and 794 will work to their most inboard positions. This may tend to
maintain the wedge members in an untwisted orientation, and may also tend
to maintain the moment arm of the restoring moment at its largest value,
both being desirable results.

[0229] When a twisting moment of the bolster relative to the side frames
is experienced, as in parallelogram deformation, all four sets of wedges
will tend to work against it. That is, the diagonally opposite pairs of
wedges in the outboard pocket of one side of the bolster and on the
inboard pocket on the other side will be compressed, and the opposite
side will be, relatively, relieved, such that a differential force will
exist. The differential force will work on a moment arm roughly equal to
the distance between the centers of the inboard and outboard pockets, or
slightly more given the gap arrangement.

[0230] In the further alternative arrangement of FIGS. 13b and 13d, a
single, stepped wedge 820 is used in place of the pair of split wedges
e.g., members 786, 788. A corresponding wedge of opposite hand is used in
the other bolster pocket.

[0231] In the further alternative embodiment of FIG. 14a, a truck bolster
830 has welded bolster pocket inserts 832 and 834 of opposite hands
welded into accommodations in its distal end. In this instance, each
bolster pocket has an inboard portion 836 and an outboard portion 838.
Inboard and outboard portions 836 and 838 share the same primary angle
α, but have secondary angles β that are of opposite hand.
Respective inboard and outboard wedges are indicated as 840 and 842, and
each seats over a vertically oriented spring 844, 846. In this case
bolster 830 is similar to bolster 680 of FIG. 11a, to the extent that the
bolster pocket is continuous--there is no land separating the inner and
outer portions of the bolster pocket. Bolster 830 is also similar to
bolster 710 of FIG. 11c, except that rather than the bolster pockets of
opposite hand being separated, they are merged without an intervening
land.

[0232] In the further alternative of FIG. 14b, split wedge pairs 848, 850
(inboard) and 852, 854 (outboard) are employed in place of the single
inboard and outboard wedges 840 and 842.

[0233] In some instances the primary angle of the wedge may be steep
enough that the thickness of section over the spring might not be overly
great. In such a circumstance the wedge may be stepped in cross section
to yield the desired thickness of section as show in the details of FIGS.
14c and 14d.

[0234] FIG. 15a shows the placement of a low friction bearing pad for
bolster 680 of FIG. 11a. It will be appreciated that such a pad can be
used at the interface between the friction damper wedges of any of the
embodiments discussed herein. In FIG. 15a, the truck bolster is
identified as item 860 and the side frame is identified as item 862. Side
frame 862 is symmetrical about the truck centerline, indicated as 864.
Side frame 862 has side frame columns 868 that locate between the inner
and outer gibs 870, 872 of truck bolster 860. The spring group is
indicated generally as 874, and has eight relatively large diameter
springs arranged in two rows, being an inboard row and an outboard row.
Each row has four springs in it. The four central springs 876, 877, 878,
879 seat directly under the bolster end 880. The end springs of each row,
881, 882, 883, 884 seat under respective friction damper wedges 885, 886,
887, 888. Consumable wear plates 889, 890 are mounted to the wide, facing
flanges 891, 892 of the side frame columns, 888. As shown in FIG. 15b,
plates 889, 890 are mounted centrally relative to the side frames,
beneath the juncture of the side frame arch 892 with the side frame
columns. The lower longitudinal member of the side frame, bearing the
spring seat, is indicated as 894.

[0235] Referring now to FIGS. 15c and 15e, bolster 860 has a pair of left
and right hand, welded-in bolster pocket assemblies 900, 902, each having
a cast steel, replaceable, welded-in wedge pocket insert 904. Insert 904
has an inboard-biased portion 906, and an outboard-biased portion 908.
Inboard end spring 882 (or 881) bears against an inboard-biased split
wedge pair 910 having members 912, 914, and outboard end spring 884 (or
883) bears against an outboard-biased split wedge pair 916 having members
918, 920. As suggested by the names, the outboard-biased wedges will tend
to seat in an outboard position as the suspension works, and the
inboard-biased wedges will tend to seat in an inboard position.

[0236] Each insert portion 906, 908 is split into a first part and a
second part for engaging, respectively, the first and second members of a
commonly biased split wedge pair. Considering pair 910, inboard leading
member 912 has an inboard planar face 924, that, in use, is intended
slidingly to contact the opposed vertically planar face of the bolster
pocket. Leading member 912 has a bearing face 926 having primary angle
α and secondary angle β. Trailing member 914 has a bearing
face 928 also having primary angle α and secondary angle β,
and, in addition, has a transition, or step, face 930 that has a primary
angle α and a tertiary angle φ.

[0237] Insert 904 has a corresponding an array of bearing surfaces having
a primary angle α, and a secondary angle β, with transition
surfaces having tertiary angle φ for mating engagement with the
corresponding surfaces of the inboard and outboard split wedge members.
As can be seen, a section taken through the bearing surface resembles a
chevron with two unequal wings in which the face of the secondary angle
β is relatively broad and shallow and the face associated with
tertiary angle φ is relatively narrow and steep.

[0238] In FIG. 15e, it can be seen that the sloped portions of split wedge
members 918, 920 extend only partially far enough to overlie a coil
spring 926. In consequence, wedge members 918 and 920 each have a base
portion 928, 930 having a fore-and-aft dimension greater than the
diameter of spring 926, and a width greater than half the diameter of
spring 926. Each of base portions 928, 930 has a downwardly proud,
roughly semi-circular boss 932 for seating in the top of the coil of
spring 926. The upwardly angled portion 934, 936 of each wedge member
918, 920 is extends upwardly of base portion 928, 930 to engage the
matingly angled portions of insert 904.

[0239] In a further alternate embodiment, the split wedges can be replaced
with stepped wedges 940 of similar compound profile, as shown In FIG. 15f
In the event that the primary wedge angle is relatively steep (i.e.,
greater than about 45 degrees when measured from the horizontal, or less
than about 45 degrees when measured from the vertical). FIG. 15g shows a
welded in insert 942 having a profile for mating engagement with the
corresponding wedge faces.

[0240] FIGS. 16a and 16b illustrate a bolster, side frame and damper
arrangement in which dampers 960, 961 are independently sprung on
horizontally acting springs 962, 963 housed in side-by-side pockets 964,
965 in the distal end of bolster 970. Although only two dampers are
shown, it will be understood that a pair of dampers faces toward each of
the opposed side frame columns. Dampers 960, 961 each include a block 968
and a consumable wear member 972, the block and wear member having male
and female indexing features 974 to maintaining their relative position.
An arrangement of this nature permits the damper force to be independent
of the compression of the springs in the main spring group. A removable
grub screw fitting 978 is provided in the spring housing to permit the
spring to be pre-loaded and held in place during installation.

[0241] FIGS. 17a, 17b and 17c show a preferred truck 970, having a bolster
972, a side frame 974, a spring group 976, and a damper arrangement 978.
The spring group has a 5×3 arrangement, with the dampers being in a
spaced arrangement generally as shown in FIG. 11c, and having a primary
damper angle that may tend to be somewhat sharper given the smaller
proportion of the total spring group that works under the dampers (i.e.,
4/15 as opposed to 4/9 in FIG. 11c.

[0242] In one embodiment of truck 970, as might preferably be used in the
location of end trucks 88, 206, 210, or 232, there may be a 5×3
spring group arrangement, the spring group including 11 coils each having
a spring rate in the range of 550-650 lb./in, and most preferably about
580 lb./in; and 4 springs (under the dampers, in a four corner
arrangement) having a spring rate in the range of 450-550 lb./in, most
preferably about 500 lb./in, for which the dampers are driven by 20-25%
of the force of the spring group, preferably about 24%. The dampers may
have a primary angle of 35-45 deg., preferably about 40 deg. In this
preferred end truck embodiment, the overall group vertical spring rate is
in the range of 8,000 to 8,500 lb./in., in particular about 8380 lb./in.

[0243] In another embodiment of truck 970, such as might preferably be
used in the location of internal truck 234, there may be a 5×3
spring group arrangement in which the spring group may include 11 outer
springs having a spring rate of about 550-650 lb./in., and most
preferably about 580 lb./in; 4 springs (under the dampers, in a four
corner arrangement) having a spring rate in the range of 550-650 lb./in,
and most preferably about 600 lb./in.; and six inner coils having a
spring rate in the range of 250-300 lb./in., most preferably about 280
lb./in. The overall spring rate for the 5×3 group is in the range
of 10,000-11,000 lb./in., and most preferably about 10,460 lb./in. The
dampers are driven by about 20-25% of the total force of the spring
group, preferably about 23%. The dampers have a primary angle in the
range of 35-35 degrees, preferably about 40 degrees.

[0244] It will be appreciated that the values and ranges given for truck
970 depend on the expected empty weight of the railcar, the expected
lading, the natural frequency range to be achieved, the amount of damping
to be achieved, and so on, and may accordingly vary from the preferred
ranges and values indicated above.

[0245] In the embodiments of FIGS. 2a, 2b, 3a, 3b, 4a and 4b, the ratio of
the dead sprung weight, WD, of the rail car unit (being the weight of the
car body plus the weight of the truck bolster) without lading to the live
load, WL, namely the maximum weight of lading, be at least 1:1. It is
advantageous that this ratio WD:WL lie in the range of 1:1 to 10:3. In
one embodiment of rail car of FIGS. 2a, 2b, 3a, 3b, 4a and 4b the ratio
can be about 1.2:1. It is more advantageous for the ratio to be at least
1.5:1, and preferable that the ratio be greater than 2:1.

[0246] The embodiments described herein have natural vertical bounce
frequencies that are less than the 4-6 Hz. range of freight cars more
generally. In addition, a softening of the suspension to 3.0 Hz would be
an improvement, yet the embodiments described herein, whether for
individual trucks or for overall car response can employ suspensions
giving less than 3.0 Hz in the unladen vertical bounce mode. That is, the
fully laden natural vertical bounce frequency for one embodiment of rail
cars of FIGS. 2a, 2b, 3a, 3b, 4a and 4b is 1.5 Hz or less, with the
unladen vertical bounce natural frequency being less than 2.0 Hz, and
advantageously less than 1.8 Hz. It is preferred that the natural
vertical bounce frequency be in the range of 1.0 Hz to 1.5 Hz. The ratio
of the unladen natural frequency to the fully laden natural frequency is
less than 1.4:1.0, advantageously less than 1.3:1.0, and even more
advantageously, less than 1.25:1.0.

[0247] In the embodiments described above, it is preferred that the spring
group be installed without the requirement for pre-compression of the
springs. However, where a higher ratio of dead sprung weight to live load
is desired, additional ballast can be added up to the limit of the truck
capacity with appropriate pre-compression of the springs. It is
advantageous for the spring rate of the spring groups be in the range of
6,400 to 10,000 lbs/in per side frame group, or 12,000 to 20,000 lbs/in
per truck in vertical bounce.

[0248] In the embodiments of FIGS. 9a, 11a, and 17a, the gibs are shown
mounted to the bolster inboard and outboard of the wear plates on the
side frame columns. In the embodiments shown herein, the clearance
between the gibs and the side plates is desirably sufficient to permit a
motion allowance of at least 3/4'' of lateral travel of the truck bolster
relative to the wheels to either side of neutral, advantageously permits
greater than 1 inch of travel to either side of neutral, and more
preferably permits travel in the range of about 1 or 11/8'' to about 15/8
or 1 9/16 inches to either side of neutral, and in one embodiment against
either the inboard or outboard stop.

[0249] In a related feature, in the embodiments of FIGS. 9a, 11a and 17a,
the side frame is mounted on bearing adapters such that the side frame
can swing transversely relative to the wheels. While the rocker geometry
may vary, the side frames shown, by themselves, have a natural frequency
when swinging of less than about 1.4 Hz, and preferably less than 1 Hz,
and advantageously about 0.6 to 0.9 Hz. Advantageously, when combined
with the lateral spring stiffness of a spring group in shear, the overall
lateral natural frequency of the truck suspension, for an unladen car,
may tend to be less than 1 Hz for small deflections, and preferably less
than 0.9 Hz.

[0250] The most preferred embodiments of this invention combine a four
cornered damper arrangement with spring groups having a relatively low
vertical spring rate, and a relatively soft response to lateral
perturbations. This may tend to give enhanced resistance to hunting, and
relatively low vertical and transverse force transmissibility through the
suspension such as may give better overall ride quality for high value
low density lading, such as automobiles, consumer electronic goods, or
other household appliances, and for fresh fruit and vegetables.

[0251] While the most preferred embodiments combine these features, they
need not all be present at one time, and various optional combinations
can be made. As such, the features of the embodiments of the various
figures may be mixed and matched, without departing from the spirit or
scope of the invention. For the purpose of avoiding redundant
description, it will be understood that the various damper configurations
can be used with spring groups of a 2×4, 3×3, 3:2:3,
3×5 or other arrangement. Similarly, although the discussion
involves trucks for rail road cars for carrying low density lading, it
applies to trucks for carrying relatively fragile high density lading
such as rolls of paper, for example, where ride quality is an important
consideration although high density lading may tend to require a stiffer
vertical response than automobiles. Further, while the improved ride
quality features of the damper and spring sets are most preferably
combined with a low slack, short travel, set of draft gear, for use in a
"No Hump" car, these features can be used in cars having conventional
slack and longer travel draft gear.

[0252] It will be understood that the features of the trucks of FIGS. 6a,
6b, 7a, 7b, 8, and 9a, 9f are provided by way of illustration, and that
the features of the various trucks can be combined in many different
permutations and combinations. That is, a 2×4 spring group could
also be used with a single wedge damper per side. Although a single wedge
damper per side arrangement is shown in FIGS. 6a and 7a, a double damper
arrangement, as shown in FIGS. 8 and 9a may tend to provide enhanced
squaring of the truck and resistance to hunting. A 3×3 or
3×5, or other arrangement spring set may be used in place of either
a 3:2:3 or 2×4 spring set, with a corresponding adjustment in
spring seat plate size and layout. Similarly, the trucks can use a wide
sideframe window, and corresponding extra long wheel base, or a smaller
window. Further, each of the trucks could employ a rocking bottom spring
seat, as in FIG. 6b, or a fixed bottom spring seat, as in FIG. 7a, 8 or
9a.

[0253] As before, the upper rocker seats are inserts, typically of a
hardened material, whose rocking, or engaging surface 480 has a radius of
curvature of about five inches, with the center of curvature (when
assembled) lying above the upper rockers (i.e., the surface is upwardly
concave).

[0254] In each of the trucks shown and described herein, for a fully laden
car type, the lateral stiffness of the sideframe acting as a pendulum is
less than the lateral stiffness of the spring group in shear. In one
embodiment, the vertical stiffness of the spring group is less than
12,000 Lbs./in, with a horizontal shear stiffness of less than 6000
Lbs./in. The pendulum has a vertical length measured (when undeflected)
from the rolling contact interface at the upper rocker seat to the bottom
spring seat of between 12 and 20 inches, preferably between 14 and 18
inches. The equivalent length Leq, may be in the range of 8 to 20
inches, depending on truck size and rocker geometry, and is preferably in
the range of 11 to 15 inches, and is most preferably between about 7 and
9 inches for 28 inch wheels (70 ton "special"), between about 81/2 and 10
inches for 33 inch wheels (70 ton), 91/2 and 12 inches for 36 inch wheels
(100 or 110 ton), and 11 and 131/2 inches for 38 inch wheels (125 ton).
Although truck 520 or 600 may be a 70 ton special, a 70 ton, 100 ton, 110
ton, or 125 ton truck, it is preferred that truck 520 or 600 be a truck
size having 33 inch diameter, or even more preferably 36 or 38 inch
diameter wheels.

[0255] In the trucks described herein according to the present invention,
Lresultant, as defined above, is greater than 10 inches, is
advantageously in the range of 15 to 25 inches, and is preferably between
18 and 22 inches, and most preferably close to about 20 inches. In one
particular embodiment it is about 19.6 inches, and in another particular
embodiment it is about 19.8 inches.

[0256] In the trucks described herein, for their fully laden design
condition which may be determined either according to the AAR limit for
70, 100, 110 or 125 ton trucks, or, where a lower intended lading is
chosen, then in proportion to the vertical sprung load yielding 2 inches
of vertical spring deflection in the spring groups, the equivalent
lateral stiffness of the sideframe, being the ratio of force to lateral
deflection measured at the bottom spring seat, is less than the
horizontal shear stiffness of the springs. The equivalent lateral
stiffness of the sideframe ksideframe is less than 6000 Lbs./in. and
preferably between about 3500 and 5500 Lbs./in., and more preferably in
the range of 3700-4100 Lbs./in. By way of an example, in one embodiment a
2×4 spring group has 8 inch diameter springs having a total
vertical stiffness of 9600 Lbs./in. per spring group and a corresponding
lateral shear stiffness kspring shear of 4800 lbs./in. The sideframe
has a rigidly mounted lower spring seat. It is used in a truck with 36
inch wheels. In another embodiment, a 3×5 group of 51/2 inch
diameter springs is used, also having a vertical stiffness of about 9600
lbs./in. in a truck with 36 inch wheels. It is intended that the vertical
spring stiffness per spring group be in the range of less than 30,000
lbs./in., that it advantageously be in the range of less than 20,000
lbs./in and that it preferably be in the range of 4,000 to 12000 lbs./in,
and most preferably be about 6000 to 10,000 lbs./in. The twisting of the
springs has a stiffness in the range of 750 to 1200 lbs./in. and a
vertical shear stiffness in the range of 3500 to 5500 lbs./in. with an
overall sideframe stiffness in the range of 2000 to 3500 lbs./in.

[0257] In the embodiments of trucks in which there is a fixed bottom
spring seat, the truck may have a portion of stiffness, attributable to
unequal compression of the springs equivalent to 600 to 1200 Lbs./in. of
lateral deflection, when the lateral deflection is measured at the bottom
of the spring seat on the sideframe. Preferably, this value is less than
1000 Lbs./in., and most preferably is less than 900 Lbs./in. The portion
of restoring force attributable to unequal compression of the springs
will tend to be greater for a light car as opposed to a fully laden car,
i.e., a car laden in such a manner that the truck is approaching its
nominal load limit, as set out in the 1997 Car and Locomotive Cyclopedia
at page 711.

[0258] The double damper arrangements shown above can also be varied to
include any of the four types of damper installation indicated at page
715 in the 1997 Car and Locomotive Cyclopedia, whose information is
incorporated herein by reference, with appropriate structural changes for
doubled dampers, with each damper being sprung on an individual spring.
That is, while inclined surface bolster pockets and inclined wedges
seated on the main springs have been shown and described, the friction
blocks could be in a horizontal, spring biased installation in a pocket
in the bolster itself, and seated on independent springs rather than the
main springs. Alternatively, it is possible to mount friction wedges in
the sideframes, in either an upward orientation or a downward
orientation.

[0259] The embodiments of trucks shown and described herein may vary in
their suitability for different types of service. Truck performance can
vary significantly based on the loading expected, the wheelbase, spring
stiffnesses, spring layout, pendulum geometry, damper layout and damper
geometry.

[0260] The principles of the present invention are not limited to auto
rack rail road cars, but apply to freight cars, more generally, including
cars for paper, auto parts, household appliances and electronics,
shipping containers, and refrigerator cars for fruit and vegetables. More
generally, they apply to three piece freight car trucks in situations
where improved ride quality is desired, typically those involving the
transport of relatively high value, low density manufactured goods.

[0261] Various embodiments of the invention have now been described in
detail. Since changes in and or additions to the above-described best
mode may be made without departing from the nature, spirit or scope of
the invention, the invention is not to be limited to those details.